The work is intended as a compact reference volume for internal combustion engines in general. with particular emphasis on diesel engines. This literature survey. lubrication. the contact pressure between ring and liner. Tamminen. miniaturised experimental work and full-scale engine testing. Central topics discussed in this work are the basic functions of the piston and the piston rings.
3
. Jaana & Sandström. hot gas erosion damage. friction. A literature survey. Recent studies include modelling. aims to shed new light on the tribological issues related to the piston assembly. the design and the materials of the components.Andersson. tribology
Abstract
The tribological considerations in the contacts formed by the piston skirt. the lubrication conditions and the influence of combustion products. mechanical and thermal loads on the rings. wear. covering over 150 references. not least indicated by the large number of articles published on this topic in recent years. Carl-Erik. piston rings and cylinder liner have attracted much attention over several decades. Piston ring tribology. the sealing action. the wear of the sliding surfaces and surface technology for wear reduction. Peter. the coefficient of friction and the friction force. blow-by leakage. VTT Tiedotteita – Research Notes 2178. exhaust emissions. Espoo 2002. Keywords Piston ring. 105 p.

Mauri Airila from Helsinki University of Technology. At the time of issue of the report. Espoo. is part of the Finnish research programme ProMotor. on the 3rd of December. by Volvo Technological Development Corporation in Sweden. which is running during the years 1999–2002. Martti Larmi from Helsinki University of Technology. 2002 The Authors
4
. Tribology of Internal Combustion Engines (in Finnish: Polttomoottorien tribologia). by the Finnish companies Fortum Oil & Gas Oy.Preface
This literature study is part of the project. Steering Group Chairman Mr Matti Säynätjoki from the National Technology Agency. Tekes Mr Erkko Fontell from Wärtsilä Corporation Mr Kalevi Salmén from Sisu Diesel Oy Mr Matti Kytö from VTT Processes Prof. and by the Technical Research Centre of Finland (VTT). Wärtsilä Corporation and Sisu Diesel Oy. Colleagues and project co-workers from VTT and Helsinki University of Technology are also acknowledged for fruitful discussions related to piston tribology. as well as the members of the project Steering Group for their support in the work. Finland. The work was financially supported by the National Technology Agency Tekes. Sisu Diesel Oy. the Internal Combustion Engine Laboratory Prof. the Laboratory of Machine Design The authors wish to thank the participating companies and institutions for their financial and technical support in the project. The project. Fortum Oil & Gas Oy and Volvo Technological Development Corporation are gratefully acknowledged. the work was overseen by a Steering Group comprising the following members: Mr Jorma Niskala from Fortum Oil & Gas Oy. Discussions with and advice from representatives of the Wärtsilä Corporation.

have attracted much attention over several decades. by limiting the flow of engine oil to the combustion chamber Good sealing capability and low blow-by for supporting the power efficiency rate Good resistance against mechano-thermal fatigue. dynamic seals support the cleanliness of a lubricant and a tribological element to be protected from external contamination. Introduction
There are two entirely diverse points of view that make dynamic seals particularly demanding in a tribological sense. caused by reciprocating contact with the piston or the linear bearing. Wear at the counter surface.1. which is an issue of increasing relevance for environmental protection and cleanliness. the counter surfaces of a dynamic seal operate under the same tribological laws as any sliding couples. Firstly. Piston rings for current internal combustion engines have to meet all the requirements of a dynamic seal for linear motion that operates under demanding thermal and chemical conditions. Seals for linear motion are particularly challenging as their counter surface smoothness. This literature survey aims to shed new light on the tribological issues related to the piston assembly. and thus contributes to suppressing wear by three-body abrasion caused by contaminant particles. miniaturised experimental work and full-scale engine testing. though with the requirements of low friction and low wear and a long service life. The work is intended as a compact reference volume for internal combustion engines in general. for ensuring a long operational lifetime Low wear of the cylinder liner. which in turn change the preconditions for the dynamic seal. low friction and system wear. or bore. Secondly. Simultaneously the seal suppresses the leakage of lubricant from the tribosystem. with particular emphasis on literature related to diesel engines. for retaining the desired surface texture of the liner Emission suppression. chemical attacks and hot erosion Reliable operation and cost effectiveness for a significantly long time
The tribological considerations in the contacts formed by the piston skirt. Recent studies include modelling. In short. not least indicated by the large number of articles published on this topic in recent years. the following requirements for piston rings can be identified: • • • • • • • Low friction. for supporting a high power efficiency rate Low wear of the ring.
8
. A common feature of all dynamic seals for linear motion is that they operate against a reciprocating counter surface. for which reason they have to be optimised in terms of sealing ability. direction of sliding. speed and lubrication tend to vary more than in the closed contact forming a seal for a rotating motion. piston rings and cylinder liner. leads to changes in the surface quality of the counter surface.

This development increased the effective pressure on the piston. when the combustion conditions became even more demanding.2. Ramsbottom and Miller were among the pioneers to investigate the behaviour of the piston rings in steam engines. i. especially designed to appropriately distribute the oil on the cylinder liner and to scrape off surplus oil to be returned to the crankcase. The oil control rings were. hence providing a higher sealing force. and the need for considering thermal expansions and clearances was smaller. to distribute and control the oil. to transfer heat. lubrication by the splashing of the rotating crankshaft into the crankcase oil surface. When fitted in a groove in a piston. when the clearances were at their maximum. in 1854. Previous piston rings had consisted of multiple pieces and with springs to provide an adequate sealing force against the cylinder bore. The first piston rings used in an engine had the sole task of sealing off the combustion chamber. constructed a single-piece. i. thus preventing the combustion gases from trailing down into the crankcase. The piston is designed for thermal expansion. Piston rings
In the early steam engines no piston rings were used. A proper lubricant film on the piston. oil control rings were introduced. This new solution enabled the use of more flexible rings.1 Main functions of piston rings
The functions of a piston ring are to seal off the combustion pressure. This made it necessary to use a sealant between the piston and the cylinder liner to allow a decrease in the clearance in cold conditions. Miller.e. with higher temperatures. and are. and to stabilise the piston. with a desired gap between the piston surface and liner wall. In the early days. the ring was pressed against the cylinder bore by its own elasticity.
2. the ring pack was lubricated solely by splash lubrication. Subsequently. introduced a modification to the Ramsbottom ring. The free diameter of the ring was 10 per cent larger than the diameter of the cylinder bore. Ramsbottom.e. piston rings and liner wall was required in order to prevent damage. The temperatures and the steam pressures were not as high as the corresponding parameters in today’s internal combustion engines. pressures and piston speeds.e. which caused stronger heat expansion of the piston material. The
9
. Increasing power demands required higher temperatures. Keeping the clearance between the piston and liner wall at a minimum considerably reduces the combustion gas flow from the combustion chamber past the piston. 2000). which conformed better to the cylinder bore (Priest and Taylor. This modification consisted of allowing the steam pressure to act on the backside of the ring. i. in 1862. metallic piston ring.

1. 2. sealing off the route down to the crankcase. The position and design of the ring pack is shown in Fig. including at least one compression ring. 2. The oil control rings used in diesel engines are two-piece assemblies and sparkignited engine oil control rings may be three-piece assemblies as well.
2. the piston skirt and rings. which relatively well isolates the combustion chamber from the crankcase.rings and the ring grooves form a labyrinth seal.1. For example. Piston rings forces are discussed in greater detail in Section 5. which usually consists of 2–5 rings. see the Figs. The piston rings support the piston and thus reduce the slapping motion of the piston. especially during cold starts where the clearance is greater than in running conditions.5b.5a and 2. In addition to the general compression rings and oil control rings there are scraper rings. 2. The ring face conforms to the liner wall and moves in the groove. at the ring gap. Furthermore.e. i. which have the tasks of both sealing and scraping off the oil from the liner wall. the piston rings prevent excess lubrication oil from moving into the combustion chamber by scraping the oil from the liner wall during the downstroke.2 Ring categories
Piston rings form a ring pack. hence easily assembled onto the piston. Some of the combustion chamber heat energy is transferred through the piston to the piston boundaries. like ring and liner conformability. see Fig. but usually comprises 2–4 compression rings and 0–3 oil control rings.
10
. fast speed four-stroke diesel engines have 2 or 3 compression rings and a single oil control ring. pre-tension of the ring. from which heat transfers to the liner wall. and gas force distribution on the ring faces. The number of rings in the ring pack depends on the engine type. Scraper rings have a beak intended for scraping off the oil. The rings are generally open at one location. The sealing ability of the ring depends on a number of factors.2.

2a. with a rectangular cross-section.
11
. preventing the combustion gases from trailing down to the crankcase.2b) brings the benefit mentioned above. 2. 2. they have a larger free diameter than the cylinder liner. The cylinder gas pressure acts on the back-side of the ring. 2. which reduces the wear rate during running-in. see Fig.1 Compression rings The compression ring acts as a gas seal between the piston and the liner wall. A tapered face profile has a good oil-scraping ability. see Fig. and the ring can be used as an oil-scraping ring as well as a compression ring. satisfactorily meet the sealing demands of ordinary running conditions and this type of compression ring is the most common one. i. The ring may have a tapered face profile in order to shorten the running-in period. The tapered face profile enables the compression gas pressure to act on the face-side as well and thus relieve the pressure against the liner wall.2.e. 2.1. Piston and piston rings. Plain compression rings. as the compression pressure is able to act on the face-side of the barrel-shaped ring and thus counteract some of the force owing to ring pre-tension. which assists the ring in conforming to the liner.2c.
2. The ring force distribution depends on the face form. The rings have a certain pretension. pressing it against the liner. The use of rings with a barrel-shaped face profile (Fig. especially on the top ring.Fig. With a rectangular face profile the force is higher than with a barrel-shaped face.

Bevelled rings can be used as compression rings.2.3. 2. 2.
Fig. see Fig.4.5b.
2.2. which causes an additional stress on the ring. The number of oil control rings in a ring pack is one or two. 2. 2. 2. see Fig. High temperature may cause the lubricant in the groove to carbonise.2 Oil control rings In addition to the task of the compression rings to seal off the combustion chamber from the crankcase.5d. In running conditions the bevelled ring is pressed flat against the liner wall owing to the gas pressure.
12
. The bevelled profile causes the ring to twist in the ring groove during engine operation. The wedge form makes the ring´s axial clearance greater at increasing radial groove clearance.
Fig.3. The appearance of the oil control ring differs from that of the compression ring. Bevelled ring edge configurations (ISO 6621-1).5a and 2. there needs to be some mechanism to distribute the oil evenly onto the liner.5c and 2. which are usually used as the second compression rings. Compression ring cross-section (ISO 6621-1). see the Figs. Normally a single oil control ring is sufficient but on occasions a second ring may be required. The wedge-type profile or (half) keystone profile is used in order to prevent the ring from seizing in the groove. 2. Scraper rings. see the Figs. can simultaneously be used as oil-scraper rings.

The scraped oil may run through the possible gap between the liner wall and the piston skirt. as well as with the lubricant. 2. With the latter alternative. if applied.1. the material should be resistant against damage even in emergency conditions. Grey cast iron is used as the main material for piston rings (Federal Mogul. even though these are the rings that control the oil film. Furthermore.
Fig. As one task of the rings is to conduct heat to the liner wall. Two-stroke spark-ignited engines. Elasticity and corrosion resistance of the ring material is required. The oil control ring is perforated by slots in the peripheral direction. the
13
. contrary to the compression rings. From a tribological point of view. good thermal conductivity is required. and therefore need no oil control rings. see Fig. 2. Compression and oil control rings (ISO 6621-1). The additional force on the oil control rings causes them to have the most extreme lubrication conditions. have the lubrication oil mixed in the fuel.4.Fig. Half keystone ring (ISO 6621-1). The ring coating. Oil control rings are not always necessary. The oil control rings may have a coil spring inserted.3 Piston ring materials and coatings
A piston ring material is chosen to meet the demands set by the running conditions. the oil is forced in front of the oil control ring. which provides a way for the excess oil to leave the ring pack area.
2. 1998). as the pre-tension of the ring is not sufficient in all instances. 2. for example.5. needs to work well together with both the ring and the liner materials. The scraped oil is collected in the oil control ring groove and transported through the piston back to the crankcase.

1989.. Mollenhauer. chromium nitride (CrN). A dense chromium carbide coating. alumina-titania (Al2O3-TiO2). tungsten carbide (WC) with metallic cobalt binder. 2000). Thin. copper and tin. metal composites. CrC-NiCr (Dufrane. thermally (plasma) sprayed with molybdenum. however coatings of this type are currently used exclusively for small series production for competition engines and selected production engines (Federal Mogul. ferro-oxides. 1997). which is used in abrasive and corrosive conditions where running conditions are severe. in addition to chromium plating. Broszeit et al. and the coating is claimed to be more wear resistant than a chromium plated or phosphated surface. MoSi2. 2001). 1999). 1992). hard coatings produced by PVD or CVD include coating compositions like titanium nitride (TiN). as some examples (Federal Mogul. 1997). Furthermore. 1998. 1992). the graphite phase can act as an oil reservoir that supplies oil at dry starts or similar conditions of oil starvation (Glaeser. Haselkorn and Kelley have investigated coatings for use in low-heat rejection engines. Surface coatings/treatments for the entire piston ring surface are based on phosphorus. Hard chromium layers can be improved by plasma spraying chromium ceramic on the ring face. nitrides. particularly when the number of layers is high (Zhuo et al. thus increasing the thermal load capacity. as a dry lubrication effect of the graphite phase of the material can occur under conditions of oil starvation.. metal-ceramic composites or ceramic composites. as a uniform coating or an inlay coating material (Mollenhauer. Piston ring surfaces are. Coatings for rings are widely used. Multilayer TiTiN coatings have been experimentally deposited onto cast-iron piston rings. Hard chrome plating is particularly relevant for the compression ring. produced by HVOF coating was found promising for piston ring applications in the work by Rastegar and Richardson (Rastegar and Richardson.
14
. 1998.grey cast iron is beneficial. Radil. Experimental work with new powder compositions for thermal spraying has included molybdenum-nickel-chromium alloys. and further chrome nitride applied by low-temperature arc vapour are coatings with properties that meet the demands in low-heat rejection engines (Haselkorn and Kelley. One example of such a coating is chromium. They conclude that high carbon iron-molybdenum blend and chrome-silica composite applied by plasma spray. 1997). chromium oxide (Cr2O3) with metallic chromium binder.

5. Silicon nitride and silicon carbide performed satisfactorily under oillubricated sliding conditions. silicon nitride (Dufrane. The cylinder liner is never ideally cylindrical. Unlubricated sliding turned out to be detrimental to the ceramics. is not always desirable. Conformability can be improved by increasing the tangential load or by decreasing the momentum of inertia. cylinder head bolt tightening and abrasion.2 for details. as higher tangential loads increase the friction. 1989). The deformation is caused by thermal and mechanical loading. 1996). This is the case especially at runningin conditions where the counterparts do not conform to each other.The possibility of using ceramic piston rings as a complement to metallic rings in advanced engine applications has been investigated. This. This requires good conformability of the rings. The ring needs to be flexible to allow rapid changes for obtaining the required shape. on the other hand. see Section 4. The piston ring cylinder liner contact is a dynamic environment.
15
.4. the varying shape of the cylinder liner in both the longitudinal and peripheral directions makes the sealing even more difficult in dynamic conditions. As mentioned above.4 Sealing ability of piston rings
Piston rings have to meet various functional demands.
2.
2. A low momentum of inertia is acquired by decreasing the wall thickness of the ring. Miniature tribotests with ceramic materials have included monolithic zirconia. the liner cross-section at the location of the piston ring changes shape during the piston motion between the TDC and BDC locations. which in turn increases the possibility of ring damage. and silicon nitride with a gradient of titanium nitride on the sliding surface (Kustas and Buchholtz. Piston rings on the other hand are manufactured nearly circular. as the piston rings are not able to fully conform to the shape of the liner. Deformation of the cylinder liner increases the oil consumption. In addition to the problems related to the static sealing capability. the cylinder liner has a non-cylindrical axial shape.1. as described in Section 2. In addition to the cylinder liner deformation experienced by the piston ring under static conditions.1 Conformability Conformability means the ability of the piston rings to conform to a deformed cylinder liner. while zirconia suffered from thermal shock cracking. sintered silicon carbide.

(b) past the front side of the piston ring at starved lubrication conditions. Measurements have shown that the twist of the piston rings affects the amount of blowby past the ring pack. The blow-by disturbs the piston and ring lubrication by affecting the oil film: combustion gases contaminate the lubricant and cause the oil to entrain in them. At high oil film pressure conditions the surface asperities deform elastically. which may result in radial collapse or axial movement of the ring (Richardson. cause high land pressure. The combustion gases flow past the piston ring at various locations: (a) at the piston ring gap. in turn. A negative twist on the second ring can cause instability of the ring.6. When the combustion gas reaches the crankcase it pollutes the lubrication oil. A positive twist on the second ring can.2. The hot blow-by combustion gases cause the piston and piston rings to overheat.4. This means that some blow-by will always have to be allowed. see Fig. 2.3 Blow-by prevention Blow-by is considered the phenomenon where combustion gases flow from the combustion chamber past the ring pack to the crankcase. Blow-by cannot be totally prevented as long as the rings have gaps and move in their grooves. This requires that the gap between the back-side of the ring and the ring groove is quite large and thus has a large gas-flow area.4.2 Counter surface effects The surface roughness of both the cylinder liner and piston ring face affects the lubricating oil film. 1996).
2. reducing the oil film thickness and forming conditions of elastohydrodynamic lubrication. which results in an increase in the blow-by. (c) or past the backside of the piston ring when the ring is not in contact with either of the ring-groove walls. The gap between the piston and liner wall is greater on the anti-thrust side of the piston than on the thrust side. The blow-by affects directly or indirectly the engine power (fuel) efficiency: the blow-by consumes some of the combustion power and increases the friction as a result of less favourable lubrication conditions.
16
. The oil film thickness is extremely thin where the peaks of the interacting surfaces form contacts.

6. Oil evaporates from the rings and the cylinder liner into the combustion chamber. and the oil is
17
.4. the ring-pack lubrication is dependent of the oil film left on the cylinder liner during the downstroke. Three different mechanisms of oil consumption occur in the ring-pack system. 2000. Combustion gas blow-by past the piston ring. The amount of oil left behind on the liner is dependent on the scraping effect of the rings (Gulwadi.Fig. De Petris and co-workers point out that reverse blow-by. Ariga. meaning gas flow from the crankcase to the combustion chamber during the exhaust phase. 1996).. 1996). oil is thrown off from the ring due to inertia. Gulwadi. In all cases the oil escapes from the cylinder with the exhaust gases. The top ring scrapes oil off the liner. During the upstroke. and gas blowing back towards the combustion chamber entrains oil from the ring pack.4 Oil consumption reduction and emission suppression Oil consumption mechanisms The oil consumption in the piston region has been investigated by means of both computer models and experimental measurements (De Petris et al. 1996. 2000). The evaporated oil trails behind or passes between the top ring and the liner during the downstroke.. 2.
2. is one of the most significant means by which the oil is transported into the combustion chamber (De Petris et al.

but also by the fuel-air mixture that flows into the crevices in the ring-pack region and remains there during combustion. 1998). see Fig. 1999). 2. directly affects the emissions of the engine.accumulated in front of the top ring.. oil is present in the ring pack area. The vortex builds up even more at the end of the exhaust stroke. Since environmental demands are becoming even stricter regarding the exhaust emissions of internal combustion engines. Thus the area where the air-fuel mixture gases can dwell is decreased. causing a significant amount of unburned hydrocarbons to leave the combustion chamber. 2. as the consumed oil. gases. the oil flow into the combustion chamber has to be reduced. This entrained oil comes from accumulated oil at the ring-end gaps and the leading and trailing edges of the rings. their optimisation is important. are wiped off from the liner wall and build up into a vortex. This optimisation is made by computer modelling. A method for decreasing the hydrocarbon emissions is to move the top ring higher up on the piston (Lacey and Stockwell. as has been presented by Tian and co-workers.7c. see Fig. among other things. Changes of the above type all increase the demands on the lubricant as it is exposed to higher temperatures and higher stresses. thus adding to the amount of hydrocarbon in the exhaust gas. The authors conclude.7b. these gases flow out of the crevices and mix with the exhaust gas. Additionally. 2000). Hydrocarbon emissions are not only caused by lubrication oil that becomes entrained in the combustion gases flowing into and out of the ring-pack area. for reducing the gas volume. This oil becomes entrained in the gases flowing to and from the combustion chamber. This accumulated amount of oil is thrown off the ring at TDC due to inertia (Gulwadi. 2. or the oil trailed into the combustion chamber from the ring pack. including hydrocarbons.
18
. see Fig.7. the clearance between the ring and the liner wall can be reduced. As the oil control rings are one of the factors that control the oil flow through the ring pack. that the lubrication between the rails and the liner is greatly affected by the twist of the ring (Tian et al.7a. Emission suppression Oil consumption reduction actions are related to actions for emission suppression. During the blow-down process. unburned hydrocarbons from the ring crevice area move into the combustion chamber as the pressure falls. from oil on the liner surface. 2000). Throughout the engine cycle. see Fig. At the beginning of the exhaust stroke. and from oil on the groove surfaces (Gulwadi. 2. During the exhaust phase.

1988). Gas flow in the combustion chamber and the inter-ring region (After Heywood. Appropriate ring designs can control the amount of oil on the liner to be sufficient after the downstroke. 2. which reduces the oil consumption.
Fig.
19
.The oil control rings scrape off surplus oil from the liner.7.

the piston rings seal the combustion chamber from the crankcase and transfers heat to the coolant. 2. 3. according to their structure.
3. Different piston types. are presented as follows: 1. Uncooled or oil spray-cooled. by their primary field of application or by their structure. which keeps the piston properly aligned within the cylinder bore. and the skirt.1 presents a single-piece nodular cast-iron piston. Important piston dimensions of four-stroke diesel engines are presented in Fig. The construction of a typical piston and ring assembly is shown in Fig. Mollenhauer. 1995.3. low-speed diesel engines. the pin support. Four classifications of piston types.1. The piston types are commonly categorised by their cooling arrangement.
20
. cast or forged monometal light-alloy pistons for highspeed automotive and small utility vehicle engines. 3.1 Piston types and geometry
There are many piston types developed according to the operating requirements of different engine types.1. (Röhrle.2 and Table 3. Uncooled or oil spray-cooled cast light-alloy pistons with ring-groove insert for high-speed heavy-duty diesel engines. Single-piece or composite pistons with a cooling gallery for high-speed heavy-duty and medium-speed diesel engines. are discussed in greater detail in Section 7. Pistons for two-stroke.1 (Section 2. The right-side of Fig. The geometry of these areas can vary significantly in compliance with the field of application. The piston skirt acts as a load-carrying surface. 2. Piston skirt and ring groove
The main task of the piston is to convert thermal energy into mechanical work.1. 1997). Furthermore. 3. The most essential areas of the piston are the piston top. and a two-piece composite piston is shown on the left-side of the Figure.1). 4. regarding the oil supply/cooling mechanism. Well-defined descriptions of different piston types can be found in the Refs. the ring belt including the top land.

Phenolic resin graphite
23
.85 g/cm3) compared to nodular cast iron (7. Lead and tin coatings are used on aluminium surfaces to achieve favourable running-in properties. One of the major advantages of forged steel is the higher fatigue strength. The lower parts are typically made of nodular cast iron (Röhrle. 42CrMo4V and X45CrSi9) with high high-temperature strength. or cast or forged steel (40Mn4. 1995). 2. The material used in the forged piston skirt is micro alloyed. are still employed today. Coatings to improve sliding characteristics. The new features of the design are a nitrided nodular cast-iron piston and a dual cooling nozzle system. A Ferrocompâ-piston design is presented in a paper by Lipp and Schmidt. The higher density of forged steel (7. high-quality yield-treated steel 38MnVS6.2 g/cm3) is compensated by minimising the draft angle in the areas where the surface stays unmachined and by optimising the structure of the piston assembly (Lipp and Schmidt. In a study by Inada. The surface coatings used for pistons can be divided into the following application fields: 1. pistons of high-grade ironbased materials. In large low-speed two-stroke engines. Metallic and graphite coatings are used for improving the sliding characteristics of aluminium-silicon piston alloys. the application fields 1 and 2 are the most interesting ones. 25 % higher compared to cast iron GG-70. Coatings to increase the knock resistance. The thickness of the metallic coatings on pistons for gasoline engines is typically from 1 to 2 µm. The trend to increase the power density of the engines sets requirements on enhancing the load-bearing capacity of the pistons. 4. a piston design for high engine output powers is presented.porous metallic parts. The composite piston consists of a forged steel crown and a forged steel skirt. This piston can be used with peak cylinder pressures above 25 MPa and is applicable in a diameter range from 160 to 640 mm. and the piston cooling efficiency can be improved by adopting a dual cooling nozzle system (Inada. Coatings to improve thermal properties. 3. In high mechanical and thermal load applications the top of the composite piston consists of high-strength materials such as nodular cast iron. In terms of tribology. Coatings to increase wear resistance. 2001). The elongation at break is more than three times higher and Young’s Modulus is approx. The conclusions drawn from the study state that the oxidation-resistance and the thermal fatigue life of a nodular cast-iron piston can be increased by applying the nitriding process. primarily nodular cast iron with a pearlitic base structure. 2001). especially on the unmachined surface.

Durga and co-workers have studied the influence of the surface characteristics on the frictional behaviour of rubbing surfaces by experiments on baseline and coated piston skirts against different cylinder liner coatings including an atomised spray coating called II-25D (Epoxy+BN+MoS2+Graphite). Since the ring groove and the flanges are part of the piston sealing system. To improve the wear resistance of the aluminium pistons in aluminium cylinders. the standard side-face finish of a piston ring has a surface quality of Rz = 4 µm or Ra = 0. In order to improve ring-groove wear
24
.
3. With the II-25D coating the power improvement was around 4. it is stated that the II-25D coating significantly reduces motoring friction torque relative to the baseline liners at high speeds. At the same time the increased flow of the hot blow-by gases interferes with the oil film between the sliding surfaces and may cause hot gas damage to the piston rings. The wear of the ring groove flanks can affect the effective geometry of the ring face against the cylinder liner (Dowson. the grooves are typically either induction hardened or chromium plated. 1993). The piston rings were standard production Fe-Mo coated cast-iron rings. are used in large pistons and on pistons in automotive gasoline and diesel engines.05).8 µm (ISO 6621-4). and cycle average 0. The results of the friction rig tests showed that the II-25D coating produces a significant reduction in friction coefficient in the entire piston speed range (boundary friction coefficient 0.07. A hard chromium layer can improve the wear resistance of the ring grooves in steel composite piston crowns (Röhrle. Thin metallic phosphate layers can improve the adhesion properties of these coatings. the piston ring belt consists of three ring grooves. The baseline piston skirt material used in the tests was 318 Aluminium Alloy and the coating applied was the atomised spray coating II-25D. since II-25D is a coating.. For comparison. To increase the wear resistance of the ring grooves in the pistons of heavy fuel oil engines.coatings. 1995). Damaged sealing surfaces lead to increased blow-by and reduced effective combustion pressure. the surfaces of the flanges have to be of very high quality. As a result of single-cylinder engine tests. The authors mention that. 1998). the low friction behaviour can be expected regardless of the substrate.5 % (Durga et al. the piston skirt surface can be coated with a wear-resistant iron or chromium layer covered with a thin tin layer for break-in.3 Piston ring groove
In a large number of piston designs. affecting the blow-by of the combustion gases and the oil consumption. The piston rings are situated in the grooves between the ring-groove flanges. from 10 to 20 µm thick.

Ring carriers are preferably made of Niresist. 1995). a hard chromium layer can be applied. an austenitic cast iron with a thermal expansion coefficient almost equal to that of aluminium (Röhrle. in high-performance diesel engines against wear. so-called ring carriers made of high-alloyed cast iron are cast-in.resistance in steel composite piston crowns. sometimes also the second.
25
. To protect the first ring groove.

4. Cylinder liner
Since the piston and the piston rings are moving in the cylinder, the cylinder liner constitutes an important tribological element as a sliding surface against the piston and piston rings.

To improve the wear resistance of an aluminium cylinder, a ceramic particulate phase can be cast-in into the aluminium liner during the manufacturing of the block. Aluminium cylinder blocks can be equipped with a cast-in or pressed-in cast-iron liner.

4.2 Plateau honing
The trend in the cylinder bore surface finishing has been directed by demands to reduce oil consumption, increase the durability and increase the resistance against wear and

26

scuffing. To limit hydrocarbon emissions and particles by reducing the oil consumption, the surface roughness of the cylinder liner should have an Ra value between 0.25 and 0.4 µm and an Rz value between 3 and 6 µm (Affenzeller, 1996). This can be achieved with optimally honed bore surfaces. There is a consensus for a surface texture of cleancut surfaces with smooth, flat plateaux and a regular, consistent arrangement of primary oil-retaining valleys (Lenthall, 1996). The surface finish of the cylinder liner has an influence on the scuffing resistance. The rise in friction, which leads to the scuffing, is associated with the polishing of the liner (Galligan, 1999b). Honing is applied for finishing the surface of cast-iron cylinder liners. The cutting marks of the honing form a pattern of diagonal valleys on the liner surface. The honing grooves, the volume and the direction of the valleys control the amount of oil available, by keeping the oil on the liner surface and by improving the spreading of the oil. Since the requirements of good sealing properties and optimal lubrication are contrary to each other, the demands on the topography of the cylinder liner are exacting. (Ohlsson, 1996, Lenthall, 1996). In slow-speed two-stroke diesel engines the liner surface is traditionally finished with a special cutting tool instead of honing in order to support the running in (Lenthall, 1996). The quality of the honed surface is affected by the bore geometry and diameter tolerances, the surface roughness of the fine bored surface, the machining operations and the number of machining stages, and the material, hardness and type of the honing stones. To obtain the specific surface, the liner is usually machined in four steps: 1. Rough boring – for the basic geometry. 2. Rough honing – for alignment. 3. Fine honing – for desired surface roughness. 4. Plateau-honing – for surface smoothing. Step 3 removes all the traces of the first two steps. Step 4, the plateau-honing, partly replaces the running-in process of the liner surface, which improves the dimensional tolerance of the cylinder, increases the engine efficiency and decreases oil consumption (Ohlsson, 1996). Many parameters have been used for characterising the plateau-honed surface: • 2D-parameters, such as Ra (mean deviation of the surface roughness), Rz (mean surface roughness), and Rmax. • Parameters describing the shape of the surface, such as skewness (Rsk) and kurtosis (Rku).

4.3 Macro form deviation of cylinder liners
In practice, it is normal that the cylinder liner is not perfectly cylindrical and of nominal bore size along its entire length. The bore distortion causes loss of conformity between the piston rings and cylinder liner. Limited piston ring follow-up performance, in particular caused by bore deformation, causes an increase on the lubricating oil consumption. The non-circularity of the cylinder bore can be described by a Fourier series (Affenzeller, 1996, Chittenden and Priest, 1993, Ma et al., 1996, 1997a):

28

Fig.R(φ ) = å ( Ai cos φ + Bi sin φ )
i =0
i =n
(4. 4. 1993). Bi i n
= radial co-ordinate = angular co-ordinate = amplitude constants = order = highest order distortion to be considered
The co-ordinate system and various Fourier orders of bore distortion are presented in Fig. 1993). Therefore. The zero and firstorder bore distortions (see Fig. Cylindrical Co-ordinate system and Fourier orders (After Chittenden and Priest. The allowed difference between maximum and minimum diameter of the cylinder bore may be from 10 to 100 times the thickness of the oil film between piston rings and the liner. the deviation from circularity within the manufacturing tolerances is likely to have a significant effect on the performance of the piston assembly.
29
. There are several reasons for the non-circularity of the cylinder bore.1)
where: R(φ) φ Ai.1.1. 4.1) are a function of the size and location tolerances of the cylinder bore (Chittenden and Priest. The cylinder liners are machined to a level of accuracy specified in terms of tolerances. 4.

by thermal expansion and by gas pressure is observed at the upper edge of the cylinder (Reipert and Voigt. 2001).The assembling of the engine can cause deformations into the cylinder liner. the tightening of the four cylinder head bolts of this engine is the reason. 1993). An axial plot of cylinder liner distortion is shown in Fig. if a fourth-order distortion (Fig. which leads to distortion of the cylinder bore (Chittenden and Priest. results in an outward deformation of the cylinder liner´s inner surface (Reipert and Voigt. the distortion of the cylinder liner can be axial.
30
.1) is a major component in a distorted cylinder bore of a particular engine.. Distortion of the cylinder liner due to the combustion pressure is significant only in highly rated diesel engines with thin-walled wet liners (Chittenden and Priest. (Chittenden and Priest. According to Ref. 4. An example of this is the tightening of the cylinder head bolts. 1993). 2001). The clamping of the cylinder liner between the cylinder head and the support at the lower end of the liner in the engine block. The magnitude of the thermal expansion is by far larger than the deformation caused by the clamping of the bolts (Reipert and Voigt. (Ma et al.2. 2001). In addition to the non-circularity of the cylinder bore. Chittenden and Priest. Inadequate cooling or over-cooling of a specific region of the cylinder may cause expansion differences around the circumference of the cylinder and along its length. The maximum deformation by clamping. 1997a. The gas pressure acts on the cylinder wall in a restricted area. 1993). Complicated ring geometries with several non-conformed regions arise from a combination of multiple-order bore distortions. 1993). The ring conformability reduces significantly with an increase in the order of bore distortion. 4.

radial and axial ring motion and ring twist. 5.1.1. Ring motion and ring twist about the ring centre affect the operation of the ring. = r/l = conrod length
Fig. The crank mechanics is shown in Fig. In an analysis of the piston ring lubrication. and the blow-by across the ring pack. The instantaneous speed of the reciprocating piston motion can be estimated with decent accuracy with the following formula (Maass and Klier. 1981):
v p ≈ rω (sin ϕ +
λ sin 2ϕ ) 2
(5.5. it is necessary to determine the velocity of the piston ring as a function of the crank angle. the wear of the ring and cylinder liner.
32
. Crank mechanism with piston location parameters. the oil film formation and the friction between the ring and the liner. Piston ring mechanics
5.1)
where: vp r ω ϕ λ l = instantaneous piston speed = crank radius = angular velocity of the crank = crank angle = conrod ratio. The primary motion of the piston rings is equal to the reciprocating piston motion.1 Piston ring kinematics and kinetics
One of the major requirements on the ring pack is related to the ring dynamics. 5.

These types of motion result from different loads acting on the ring. (Ejakov et al. 1995). and friction loads from the sliding contact between the ring and cylinder liner. below and behind the ring produces resultant forces on the ring section (Dowson. 1993). oil film damping loads. 1995). change proportionally to the square of the engine speed (see the Figs. the sliding friction forces and the contact pressure between the ring and the liner cause normal and tangential forces on the ring face. ring lift. 5. 5. Loads of this kind are inertia loads arising from the piston acceleration and deceleration.. 5. Fig. which causes a non-uniform distribution of the contact pressure between the cylinder liner and the piston ring face and can thus lead to increased blow-by and oil consumption (Dowson. and ring twist.2 Piston ring forces and moments
The piston ring secondary motions can be divided into piston ring motion in the transverse direction. 5.
36
. The effect of the clearance between the cylinder liner and the piston on the piston and piston ring motion and to the ring forces is presented in Section 5. The forces acting on the ring are presented in Fig.4) (Röhrle.2 and 5. The gas pressure above. Forces acting on the piston ring (After Ejakov et al. The inertia forces acting on the piston rings. 1999).7.7. as well as those acting on the other reciprocating crank mechanism components.
Fig.6. loads owing to the pressure difference across the ring.. The shearing of the lubricating film. 1993). The elastic distortion of the piston and liner can affect the effective geometry of the ring face and cylinder liner contact.5. piston ring rotation. 1999). The side loading of the piston against the cylinder wall is a result of the articulated joint of the connecting rod (Röhrle.1.

1995). This pressure presents only about 1 % or less of the peak gas pressure (Dowson. Furthermore. This arrangement is applied in order to avoid piston-generated noise or to reduce the thermal load on the ring grooves (Röhrle. the theoretical contact pressure used in the calculation of the tangential forces of rectangular and half keystone rings made of steel is approximately 0. In the conclusions of their work.
5. and crank offset from the bore centreline. 1993). The results presented predict that. the pressure is approximately 1 N/mm2. 1995).6626 (ISO 6621-4). on the mechanical efficiency and engine noise. According to work by Dowson. the contact situation will be worse and the friction losses will increase if the piston pin offset is positioned towards the minor thrust side of the piston (Chittenden and Priest.3 Ring contact pressure
5. the authors state that the piston pin offset is the most sensitive parameter producing considerable variations in kinetic energy loss and mechanical efficiency. centre of gravity. 1993).. For example. generally.The piston pin is often offset from the piston centreline.3. According to their predictions. Haddad and Tjan have used a computer program to investigate the influence of the offset of a piston pin. The spring force of the compression rings is lower than that of the oil ring (Dowson. The contact pressures arising from ring compression are defined for undamaged piston rings. Chittenden and Priest have presented the same kind of results. 1993). the loading caused by the elastic spring force on rings with gas pressure acting on the rear of the ring is typically 104–105 Pa. and the mechanical efficiency can be increased by setting the piston pin offset to the thrust side of the piston centre (Haddad and Tjan.
37
.1 Contact pressure from ring compression
The tangential force of the piston rings depends on the piston ring type and the class of nominal contact pressure. the kinetic energy loss decreases when the piston pin offset is set to the thrust side of the piston and the mechanical efficiency increases when the piston pin offset is set to the minor thrust side of the piston. In the case of coil-spring-loaded oil control rings. The situation changes significantly in the case of a ring breakage.. The values of nominal contact pressures and specific tangential forces for various piston rings are tabulated in the standards ISO 6621.19 N/mm2. the kinetic energy loss can be reduced.

The cylinder bore deformation is discussed in more detail in Section 4. the gas pressures are of significance only for a small proportion of the engine cycle (Chittenden and Priest..3. Tian has studied dynamic behaviour of piston rings.3 Additional loads from the deformation of cylinder bore and ring
While the ring and/or cylinder bore deformation leads to non-uniform contact pressures. the ring must generate greater hydrodynamic pressures in the area where the bore is closest to the ring (Ma et al.
5. 1997a). In order to support the ring loading. He states that using a symmetrical barrel face. different sections of the face surface of the ring alternate to form the ring/liner contact. wear and the surface quality of the cylinder bore together with the surface quality of the piston ring. which causes non-uniform contact pressures and reduced conformity. the ring fails to conform to the bore.3. In ring twisting.4 Non-uniform pressure distribution due to ring twist
The ring twist affects the access of the gas pressure flow behind and between the piston rings.2 Contact pressure arising from gas pressure
The gas pressure behind the first compression ring varies according to the cylinder pressure. The gas pressure behind the oil ring stays through the whole work cycle almost equal to the pressure on the crank chamber. the contact pressure and thus the conformity is greatly increased.5. On the other hand. the groove´s downward tilt angle that gives uniform average contact pressure along the radial direction is less than in using an offset parallel face (Tian. The gas pressure behind the second compression ring is already significantly lower than the pressure behind the first compression ring. 1993). With a gas pressure acting on the piston ring.3. The ring twist can be affected by the ring face profile. the local pressures can momentarily rise to considerably higher values than the mean contact pressure. 2002). The contact and surface pressures are affected on a micro-scale of the honing. In a region of larger distortion. and this leads to a non-uniform contact pressure.
5.2.
38
.

Fig. 1998). The temperature of the piston top land is generally higher than that of the piston rings. The high combustion temperature gradients on the piston top cause the hardness of the piston alloy to drop (Röhrle. 1995). 1995). The mechanical loads are superimposed on thermal stresses. The characteristics of the lubricant in response to these conditions influence on the extent of the carbon deposit formation. the
39
.6.1. and the yield strength and shear strength of the material decrease as the temperature increases.. 1991). 6.1 presents a schematic overview of the operation temperatures of diesel and gasoline automotive engines at full load (Röhrle. 1995). including the top land and the first piston ring. High heat flux causes problems such as thermal stress and deterioration of the lubricating oil film (Liu and Reitz. Figure 6. The transient peak temperatures of the gas in the combustion chamber during the combustion in a diesel engine can rise to approximately 2 500°C (Saad et al. Thermal loads
6. is directly thermally exposed to the combustion gases and thus subjected to transient temperature variations. Schematic overview of operation temperatures of automotive engines at full load (After Röhrle. Thereby. In the top land area the lubricant is subject to the highest loads with respect to its thermal and oxidative stability.1 Macro-scale thermal loads
The ring belt area.

6.1 Frictional heating in unlubricated sliding contacts
The rise in temperature of a sliding surface due to frictional heat generation depends on the frictional power (Pµ) and the interaction of several factors.2.
40
.2 Micro-scale thermal loads
The temperature on a sliding surface is in most cases higher than the bulk temperature owing to the frictional work that takes place at the sliding surface. Hence an increase in temperature is a natural consequence of the frictional work. 1991). bore polishing and scuffing (Saad et al. According to the first law of thermodynamics. FN the normal force. Frictional heating appears as the heating of fluids due to viscous flow between surfaces in relative motion.temperatures indirectly contribute to wear.
6. In addition to the above-mentioned features. while heat transfer by radiation is low. 2000). and the cooling provided by a fluid lubricant. that thermal energy cannot be completely converted into mechanical work. From 40 to 70 % of the total heat flow into the top of the piston is transferred by the piston ring belt area and the cylinder wall surface into the coolant (Röhrle. and µ the coefficient of friction. the heat expansion and heat-related distortion of the piston top land. The deposit formation results in an increase in the radiation absorption. The frictional power as such is defined as Pµ = Fµ × v = FN × µ × v (6. heat and work are two mutually convertible forms of energy. v the sliding or rolling velocity. 1995). the specific heat of the material. the temperature and volume of the surrounding material. however. and as the heating of solid surfaces in sliding or rolling contact. It is worth noticing. as is stated in the second law of thermodynamics (Matthews.1)
where Fµ is the friction force. which expresses the principle of conservation of energy. which primarily takes place by convection. the thermal conductivity of the material. piston rings and cylinder liner cause potential loss of conformity between the piston rings and cylinder liner. such as the real area of the sliding contact.. The local temperature gradients between the combustion gases and the surfaces increase the heat transfer.

the flash temperatures are significantly higher than the bulk temperature and the bulk surface temperature (Kong and Ashby. According to early work by Bowden and Tabor. 1990. 1991). may rise to the melting temperature of one of the two materials that has the lower melting temperature (Bowden and Tabor. see Fig. at a depth of a few ten of µm below the surface. while it is less sensitive to the load applied (Kong and Ashby. 6.2. or hot spots. and subsequent versions). however the principle can be adopted on boundary lubricated or starved sliding contacts. The flash temperature increases with an increase in the sliding velocity. depending on the depth of influence and the surface area proportion involved: Firstly. The above principles for the temperature distribution and levels in a sliding contact have been used for developing the T-maps software. an increase occurs in the bulk surface temperature of the material. 1991). Owing to the minor volumes of material involved in the sliding contacts. 1991).The effect of the frictional work on the temperature of the sliding surfaces can be divided into two parts. The original T-maps software deals with unlubricated contacts. Secondly. the frictional work causes local or flash temperatures at the surface asperities where the sliding contacts actually take place. which is intended for the analysis of simple sliding contacts (Ashby et al.
41
. The bulk surface temperature increases with an increase in load and velocity (Kong and Ashby. 1971). The bulk surface temperature can be approximated as being uniform over the entire sliding surface. the flash temperatures.

for a sliding contact as follows: Fη / A= η × v / h (6.T7 branching downwards from the centre of the figure) increase rapidly with increases in the sliding velocity.. 6.T7 branching to the right from the centre of the figure) require both a higher velocity and a higher load. The graph illustrates how the flash temperatures (temperature contours T1.2 Lubricant heating owing to viscous work
The temperature rise in a fluid lubricant owing to viscous work depends on the interaction of several factors like the specific heat of the lubricant and the flow rate of the fluid. which is described in the Reference (Ashby et al.2)
42
. Graph produced using T-maps 2.2..0 software. while the corresponding bulk surface temperatures (temperature contours T1..
6. Example of a map of calculated flash temperatures and bulk surface temperatures versus load (vertical axis) and sliding velocity (horizontal axis) for lowcarbon steel sliding against itself without lubrication.Fig. 1990)..2.. The viscous work depends on the dynamic viscosity and the velocity gradient between the stationary and moving surfaces.

The frictional power owing to the viscous loss is obtained as the product of the viscous force Fη and the sliding velocity v. η the dynamic viscosity of the fluid.
43
. The temperature increase owing to viscous work can normally be ignored. as it is significantly smaller than the effect of frictional work in an unlubricated or boundary lubricated sliding contact. A is the surface area of the sliding contact. v the sliding velocity and h the oil film thickness. Variations in the viscosity owing to temperature and pressure effects make the frictional power expression more complex.where Fη is the viscous force opposing the motion.

or pistons with no cooling or oil-spray cooling.1. or oil-cooled pistons with cooling system. The oil is usually supplied to the piston through the connecting rod.1 displays a so-called shaker piston. is mostly used in highspeed engines. The second group. although not all pistons require cooling. includes pistons with cooling channels and drillings.1 Oil supply
7. cooled with an oil spray. The bottom of the piston is. Figure 7.7. if required.
Fig. and crosshead engine pistons. Pistons are divided into groups as follows: pistons with no cooling or oil-spray cooling. from the main bearings.1 Supply mechanisms and oil quantity
Piston types Piston lubrication is in most cases connected to the cooling of the piston.
44
. where the upper bearing of the connecting rod has a hole from which oil spurts up into the piston. The oil is dispersed to the opposite sides of the piston´s inner space as the connecting rod oscillates during the cycle. Piston ring and skirt lubrication
7.1. oil-cooled pistons with a cooling system. 7. The first group. "Shaker"-piston.

the piston was found to tilt towards the thrust side. Thirouard and co-workers further observed that there were three possible mechanisms of oil transport to the second land: (1) The oil can be scraped to the second land in the top ring down-scraping (downstroke) or second ring up-scraping (up-stroke). oil flowed to the crown land whenever oil was present on the second land (Thirouard et al.
45
. One solution for oil supply to the piston is that the oil is fed from the main bearing or bearings to the crankshaft and further on to the connecting rod and the piston. as the amount of oil supplied this way is usually sufficient. Oil accumulation on the crown land has been investigated by Thirouard and co-workers. A second mechanism of oil transport to the crown land was also observed. The oil is further needed in the ring groove for preventing the ring from sticking to the groove. directly or indirectly. the up-scraping is impossible. 1998). The up-scraped oil was transported to the crown land. 1998). Oil is supplied to the piston and piston rings from the crankcase.. (3) The oil can be carried through the gaps in the ring pack by gases. high-speed engines use splash lubrication. Oil transport mechanisms Oil transport mechanisms have been studied by a two-dimensional laser-induced fluorescence technique. Two oil flow mechanisms on the second land were observed: (a) oil flow by inertia in the axial direction and (b) oil being dragged by the gas flow in the circumferential direction.. the oil acts as a heat carrier that transports heat from the piston and the ring-liner interface. The oil supply method usually depends on the size of the engine and on the required amount of oil.Oil supply Oil is needed at the piston ring and liner interface to provide hydrodynamic or mixed lubrication in order to reduce friction and prevent seizure. In addition to the lubricating aspect. The tilting piston and the ring twist caused the ring upper corner to start scraping oil from the liner wall. With a tapered profile on the second ring. The top ring up-scraping of oil was observed at engine speeds above 1 600 r/min. (2) The oil can flow through the two upper ring grooves. which flows either towards or from the combustion chamber (Thirouard et al. Smaller. The system enables studies on the oil distribution on piston surfaces and between the rings and the liner. Larger engines that need a higher amount of oil need to have the oil supplied up into the piston. The accumulation of oil was most probably caused by the top-ring up-scraping. as there was no hydrodynamic pressure supporting the radial forces. Oil flows into the ring groove and is pumped out of it as a result of radial ring movement in the groove. Owing to increased gas pressure.

or on the condition of the oil. contaminate the oil. The oil change intervals can be based on running hours. 2000). such as soot and wear particles.2 Oil quality
Oil degrades due to age and contamination. Undesired polishing of cylinder liners in diesel engines during operation.3 Contaminations in the oil
Oil exchange interval aspects The reason for engine oil changes is that the oil is degraded in terms of viscosity and oxidation. In the ringpack area. The additives vary depending on the operational environment for which the oil is designed. particularly in the case of smaller engines. This disadvantage is normally maintained by replacing the contaminated oil with a new batch. viscosity index modifiers. who conclude that friction in fully flooded conditions with MoDTC is clearly lower than with non-friction-modified oils.7. The most common additives are boundary friction reducers. The contaminants in the engine oil build up over time.
7. this occurs especially in the ring and land regions. The degradation is particularly caused by high temperature and blow-by gases. and that solid or liquid contaminations become mixed or dissolved into the oil. as determined
46
.1. 1998). Engine oils with and without friction-reducing additives have been investigated by Glidewell and Korcek. With age. and anti-wear additives such as ZDDP. may occur if corrosive species and small abrasive particles are present in the lubricating oil (Godfrey.. Molybdenum dialkylthiocarbamate (MoDTC) is a base oil additive that reduces the boundary friction in a surface contact. the friction-reducing effect of MoDTC seems to degrade (Glidewell and Korcek. 2000).1. The additives in the base oil fall apart and combustion particles. Friction modifiers A lubricant consists of base oil and additives. Recent investigations by Bijwe and co-workers have determined the necessity for crankcase oil drainage at set intervals (Bijwe et al. In starvation conditions the friction with the MoDTC-modified oil may decrease to become equal to that of non-friction modified oils. The presence of contaminations in engine oil is generally undesired. as solid contamination particles potentially cause abrasive wear and liquid contaminations may cause corrosive attacks. commonly called bore polishing. tribochemical wear and viscosity changes.

Crankcase oil that is contaminated by abrasive particles leads to wear of the piston skirt below the piston rings. 1998). as this may have a strong influence on how representative the sample is in regard to the condition of the engine. Bore polishing is another undesired effect of small abrasive particles in the lubricating oil (Godfrey.. The use of oil filters with an appropriate nominal retention rate is an effective tool for suppressing the
47
.. Solid contamination particles The detrimental effect of solid particles in the engine oil is particularly obvious with softer materials. chromium. if particles of sufficient size are present in relevant concentrations in the engine oil. provided the pressure is sufficient for allowing sampling. Hunt. while larger particles can cause scratches in the bore. Except for the oil oxidation products and the external contaminants (presented below) that limit the useful life of an engine oil. which is indicated by a matt appearance of the surface below the rings. particularly on the thrust sides of the piston skirt. the return line can be used for sampling.by oil sample analyses or by the response of a particle sensor (Chambers et al. tin and nickel alloys (Macián et al. 2000. 2002). and wear particles consisting of ferrous. Datoo and Fox. 1993) in the oil circuit or tank.. In dry sump engines. like piston skirts and journal bearings. 1988. 1998) give reasons for oil changes. 2000).. viscosity changes and additive depletion (Glidewell and Korcek. aluminium. Jiang and Wang. When establishing an oil analysis programme. lead. or a vacuum pump can be employed to assist the sampling (Fitch and Troyer. Piston rings do suffer from abrasive wear. because the same additives are known to contribute to the antioxidant properties of the lubricating oil (Korcek et al. 2001). particular notice should be made to the fact that the oil at the piston rings is far more contaminated than the oil in the engine sump or oil tank (Fox et al. 1997. copper. there is a higher probability for larger wear particles than during the stationary operation of the engine (Kaisheng et al. 1998). The origin of the contamination particles may be soot from combustion. ingested dust in the shape of silica dust and similar minerals. 2001). the location and time for the sampling of used engine oil should be carefully considered. and in highly loaded contacts like the cam-follower contact. 1998). which is powerful for reducing friction. The MoDTC/ZnDTP additive system.. During the running-in period of an engine. In this context. can be consumed by oxidation. by which the remaining concentrations of zincdialkylthiophosphate (ZDDP) and phenol/aminic antioxidants can be easily monitored (Jefferies and Ameye. In wet sump engines a special sampling point should be introduced between the oil pump and the oil filter. A recent addition to the analyses for evaluating the remaining useful life of an engine oil is found in the RULER™ (Remaining Useful Life Evaluation Routine) voltammetric analyses.

particle concentration in engine oil (Jiang and Wang. Oil filters in practical applications can have nominal retention rates in the order of 15. Absolute limits for the types. a method that is suitable for the determination of the presence of various organic compounds including reaction products (McClelland and Jones.. 2000). The effect of the acidic contaminants can be diminished by neutralisation. for instance the Karl Fischer method (DIN 51 777). 1998. 1995). The main categories of solid particles in crankcase oils are the carbon. in all lubricated systems the particle size should remain well below the oil film thickness in any lubricated mechanism. As a rule of thumb. and this very well applies to engines. A feature in common to all liquid contamination of lubricating oils is that they either affect the viscosity of the oil. combustion products. 2001). or wear. 1995). D664 and D974 methods (Dong et al. This is a natural consequence of
48
. The total acid number (TAN) or the corresponding total base number (TBN).the fuel residues. Exhaust gas re-circulation without soot filters has been found to increase the level of carbon particles in the lubricating oil. or combustion. metallic wear particles and oxidation particles cause less abrasive wear than external silica dust particles of higher concentration (Truhan et al. most liquid or dissolved contaminants are difficult to mechanically separate from lubricating oil. Except for water.. soot. 1993). The content of metallic particles can be quantitatively determined by atomic absorption spectrometry (ICP-AES). Carboxylic acid and other reaction products resulting from oil oxidation can be identified and quantified by means of Fourier transform infra-red spectroscopy (FTIR). can be determined with the ASTM D663... for example (Hunt.20 µm. According to experiments by Truhan and co-workers. The carbon particles can be determined by means of Fourier transform infra-red spectroscopy (FTIR). particles. size groups and concentration of contaminant particles in lubricant oils cannot be established for engines in general. as the limits are individual for each specific type of engine. Jones and Eleftherakis. as an abrasive medium. Liquid or dissolved contaminations and oil oxidation products Liquid contamination of oil can occur in a variety of ways . is more contaminated than the bulk oil volume of the engine. The water content can be determined by. condensed water or lubricant oxidation products as a few examples. or/and cause corrosion of the lubricated surfaces. or an overbased oil can be used in applications where acidic combustion products can be prospected. which represents the total effect of the acidic (and base) species that are present in the oil. The minute oil volume that is entrapped in the ring pack of a piston. particles and the metallic.

The poor quality of the oil in the inter-ring region should be taken into consideration when assessing the tribological conditions under which the piston rings operate. Gamble et al. viscosity and counter surface effects
The ring-pack area experiences different kinds of lubrication regimes owing to local speed. Geist and Barrow. (Fox. 2001). The desired lubrication conditions are the fully flooded ones. When further evolved. coke ring
The formation of carbon deposits on the piston top land can lead to the build-up of a sliding shoe. 2001. Work by Fox and co-workers has experimentally verified that the composition of the inter-ring oil is strongly different from the oil in circulation. Their findings show that. which calls for less oil at the cylinder liner and a longer residence time for the oil in the ring pack. Moritani and Nozawa. and improper lubrication of the ring pack (Fox. Fox et al.2 Lubrication regimes. 1999. and to variations in the oil supply. This is. The situation has been made more severe by the tightened exhaust emission legislation. the lubricant is subjected to substantial dilution by fuel and condensed water. which results in a destabilisation of the lubricant into several liquid and solid phases. Datoo and Fox.. changes in the viscosity properties. speed.4 Coke formation. reduction in anti-oxidancy. 1996.. Experiences of using anti-polishing rings in large-bore diesel engines are presented in the Reference (Demmerle et al. as the wear of the surface is negligible in this regime. which forms a contact with the cylinder liner thereby causing bore polishing.
7. 2001).. in particular during start-up and warming-up of the engine. load. nearly impossible to achieve with today’s engine power demands. water vapour and other combustion products in the combustion chamber of the engine. mostly mixed and boundary lubrication
49
. the bore polishing leads to a local liner wear and increased lubricating oil consumption. 2002). however. High combustion pressures and piston speeds exceed the optimal ones in terms of lubrication.1. Therefore.
7. load and surface roughness variations. 1996.the presence of fuel. 2001). reduction in the base number. The anti-polishing ring limits the thickness of the carbon deposits on the piston top land and thus restrains the contact between the liner and the piston top land (Amoser. In some engine applications the risk of bore polishing owing to the formation of carbon deposits is prevented by providing the upper part of the liner with an anti-polishing ring. 1997.

These conditions typically occur when the piston ring sliding velocity is high and a low pressure acts on the back-side of the ring. For example. while the rest of the load is carried by the surface asperities and gas forces. This occurs in the vicinity of the dead centres. or texture. Suggestions are made that a slight increase in friction. as are the conditions during piston mid-stroke. on the friction occurs under conditions of mixed and boundary lubrication. The oil film thickness of the top ring is hardly at all influenced by a change in the surface pattern and therefore the change in the surface friction is negligible. The load is carried solely by the oil film. The different lubrication regimes are briefly described below: • Fully flooded ring lubrication: The oil film covers the whole ring surface area. which could lead to oil starvation and thus friction increase. could be partially caused by high-speed shear. a high combustion pressure may disrupt the oil film. Viscosity effects The effect of the oil viscosity on the frictional behaviour of piston rings has been investigated by Durga and co-workers (Durga et al. • Starved ring lubrication: Oil availability on the liner is at its minimum. Surface roughness and surface pattern effects Surface roughness and textures have a considerable effect on the ring-pack friction. Higher friction values occur at higher viscosity. show differences in friction depending on the direction of the surface texture. • Partially flooded ring lubrication: Only a part of the ring is lubricated with oil. In this area the load is carried by the oil film. A longitudinal pattern gives the highest friction and a transverse groove pattern gives the lowest friction. The second ring and the oil control rings. Oil transport to the ring in mixed lubrication is more or less insufficient. Sui and Ariga have investigated the influence of surface patterns on oil film thickness and ring friction.. which is observed at mid-stroke of the piston motion. The rings experience a very high contact pressure at mid-stroke. after which the speed is too high for the oil film to withstand. 1998). Sui and Ariga
50
. The oil viscosity affects friction values under conditions of pure hydrodynamic lubrication when the rings are fully flooded. Increasing speed increases the oil film thickness to certain level. where the speed is lower and/or the pressure on the backside of the ring is high. where contact between sliding surfaces occurs. on the other hand. The oil is forced away from the ring surface area owing to insufficient oil supply and/or a strong gas pressure gradient over the ring.occurs in the ring-pack area. The influence of the surface pattern.

Hence.e.2. 1994). These assumptions cause the effective width of the ring face to be 20–30 % of the whole ring width. Full hydrodynamic lubrication requires that the ring area is flooded. mixed lubrication occurs. mixed and pure boundary lubrication. roughness orientation optimisation should be considered (Sui and Ariga. using this model. The equation can be solved for pressure distribution. Near the dead centres. 1993). Hu and co-workers investigated surface roughness and oil flow factors. The modes are pure hydrodynamic. there is always enough of lubrication oil between the ring and the liner wall in order to prevent surface contact. Different lubrication models for computer simulation have been developed accordingly.conclude that. viscosity. Full hydrodynamic lubrication occurs when there is no surface asperity contact. however. Lubrication conditions at mid-stroke are normally hydrodynamic. The Reynolds equation includes parameters of the geometry.e. As a result. as long as there is a sufficient amount of oil available at the leading edge of the ring.1 Oil film thickness calculations
There are many theoretical models of piston ring lubrication. load capacity. the ring face works under flooded conditions at the vicinity of the dead centres (Han and Lee. Oil film thickness calculations consider various types of lubricant modes.
7.. This is possible in the mid-stroke area where the relative surface velocity is at its highest. i. Oil film thickness fluctuations are caused by improper ring design. additionally elastohydrodynamic lubrication is on occasion considered as an extension of pure hydrodynamic lubrication. they showed that the contact pattern and the distribution of the oil film thickness between the cylinder liner and the piston ring are not exactly symmetric. i. Honing the cylinder liner with crosshatches has proved to increase the hydrodynamic action on the ring pack lubrication (Michail and Barber. 1998). They modelled the ring deflection and the contact load. Ma and co-workers have compared two different oil availability
51
. 1995). static distortion of the cylinder liner and dynamic load on the piston ring (Hu et al. the ring is partially or totally starved. in situations where the surface roughness cannot be reduced to an optimal level. The authors have developed a new model where the inlet region is in a starved condition and the outlet region has an open-end assumption. In common for almost all of these models is that they are based on the Reynolds equation. Experimental investigations by Han and Lee have shown that the ring face is not fully lubricated. pressure and surface velocities. friction force and oil flow. the Reynolds boundary condition cannot be used in this case. Still.

any circumferential variations are omitted by this approach. The authors conclude that only approximately 10–40 % of the ring face is covered by an oil film. and the elasticity of the surfaces interacts with the lubricant or under the pressure of the lubricant. the surfaces start to elastically deform according to the pressure. the surfaces approach each other more than the clearance would allow. The oil film boundary value for the mixed lubrication model can be determined in many ways but it always depends on the roughness of the surfaces involved. When the relative surface velocity decreases. where the piston rings are considered to have a sufficient amount of oil. 5. The ring-groove clearances are. Other features are too complicated to implement. A three-dimensional approach. Ring-pack simulation models include a variety of features that an actual ring/liner contact comprises.6. The oil film thickness has decreased to such a low value that the oil film only provides lubrication between the asperities but the load is carried by the surface peaks and not by the oil film. Boundary lubrication occurs when the surface contact is continuous. Considering the two-dimensional approach. on the other hand. where the pressure is high. consumes a considerable amount of computational processing time. the oil film separating the surfaces decreases in thickness. where the oil film thickness of the preceding ring is considered as the available oil film thickness for the trailing ring. namely a fully flooded model and a flow-continuity model. in comparison with a twodimensional approach.. 1997a). rather than a fully flooded model (Ma et al. Elastohydrodynamic lubrication of piston rings occurs in all internal combustion engines during the highly loaded expansion stroke. For piston tilting. Therefore a flow-continuity model should be used. An oil film model should take into consideration factors such as
52
. different on the thrust and anti-thrust sides because the piston is forced against the thrust side of the liner. In hydrodynamic lubrication. The ring/liner wall is usually modelled in two dimensions. thus considering the contact to be uniform throughout the circumference. Almost every hydrodynamic lubrication model iterates the oil film pressure in order to establish an equilibrium of the pressure value. This is not the case in actual situations. During this elastohydrodynamic lubrication regime. and others make the required processor time very long. Some features are excluded from certain models. as the piston experiences different normal forces on the thrust and the anti-thrust sides. The fully flooded model comprises a model. while the flow-continuity model comprises a model. as they are not considered important for the purpose for which the model has been developed. Asperity contact will occur if the oil film becomes thin enough. for example.models. see Fig.

. while more sophisticated lubrication models include the shear-thinning characteristics of the lubricant (Tian et al. engine speed. liner and land temperatures and surface texture (Tian et al. gas pressures. Richardson and Borman have added a varying viscosity model in their lubrication model (Richardson and Borman. not even this approach fully corresponds to reality. A model. The hydrodynamic pressure is solved from the lubrication equations. however. The measured oil film thickness is greater than the calculated.. in order to keep the model size within reasonable limits. The reason is presumed to be additional oil transported from the piston skirt. the running face profiles of the ring. Akalin and Newaz have developed an axi-symmetric. The film thickness decreases when the ring width is reduced (Seki et al. In oil film thickness calculation certain assumptions must be made.oil availability in the ring pack. as the mixed lubrication regime is dominant in this part of the stroke. and running speed are the most important parameters. However. 2001). Models differ by the way in which the oil supply to the ring/liner contact is modelled. The results show that the temperature. 1992). as they affect the lubrication regime the most. 2000). The lubricant fluid is usually considered Newtonian. 1996). 1996). Some lubrication models assume viscosity to be constant between the ring face and liner wall. Oil film thickness at the oil control ring increases when the ring tangential tension is reduced. The effect of the normal load on the friction coefficient during mixed lubrication was low. mixed lubrication model in order to simulate the piston ring and cylinder liner frictional contact. radial tension of the ring. as the oil film is under additional influence of the land environment between the rings. 1992). The oil film in front of the ring can be considered as being of constant thickness. The authors have compared the results obtained by simulation to those obtained in a test bench and found that the friction results correlate well (Akalin and Newaz. The friction coefficient does. The oil pressure in front of and
53
. but the pressure under boundary conditions still needs to be specified.. These factors prevent the oil ring from following the cylinder liner and thus causes an increase in the oil film thickness (Richardson and Borman. Oil film thickness calculations and measurements by Richardson and Borman have shown that the oil film thickness of the oil control ring differs significantly from a theoretical value at the beginning of the downstroke. hydrodynamic. Their simulation results show that the hydrodynamic lubrication regime occurs during the main part of the stroke. in which the oil film thickness trailing the previous ring is considered as the input oil film thickness for the following ring is more realistic. lubricant properties. and piston slapmotion. surface roughness. show an increase at the top and bottom dead centres.

mass-conserving algorithms are used in order to take into account the effect of cavitation zones on the oil availability. such as the engine speed. Some models are criticised for causing inaccurate results owing to what has been assumed. It is practical to study the influence of individual parameters.trailing the ring is assumed to be equal to the gas pressure on the respective side of the ring. which in turn is critical for oil consumption. viscosity of the oil. Loenne and Ziemba have suggested that the bore distortion could be described by terms of a Fourier series (Loenne. and the profile of the piston ring.1. on the lubricating oil film by using test rigs and motored engine tests. and this is designated as the cavitation zone. The bore distortion directly affects the piston/cylinder liner blow-by. There are several ways of attacking this boundary condition in the lubrication models (Richardson and Borman.2. The Sommerfeld condition allows both positive and negative pressure values. 1997b). Bore distortion in lubrication models Bore distortion plays a significant role in the conformability of the piston ring. piston friction loss and seizure. it is essential to include bore distortion in a computer model. At the top dead centre. Therefore. 1992).
54
. On the front side of the ring.2. Oil film simulation boundary conditions Various boundary conditions in the lubrication simulation are used. is so high that even larger bore distortions are corrected by the top ring conformability. Friction simulation is discussed in greater detail in Section 8. 1988). The HalfSommerfeld condition sets all negative pressure values to zero. the remaining oil on the liner either passes under the ring or accumulates in front of the ring. This suggests that bore distortion might not lead to excessive blow-by. and by this the pressure acting on the back of the top ring. who assume that the distortion only occurs in the circumferential direction and not in the axial direction of the cylinder. Furthermore. Loenne and Ziemba.2 Oil film thickness measurements
The lubricating oil film on each ring of the piston is closely related to the oil consumption. emissions and a sufficient lubricant supply. the combustion pressure.. The ring´s front edge will no longer have an oil film between it and the liner wall. 1972.
7. This approach of bore distortion has also been used by Ma and co-workers. This means that the oil film supporting the ring has less load support than a flooded ring. They additionally point out that bore distortion in some way may be desirable because it can reduce the friction loss of the ring pack (Ma et al. the ring becomes starved. If the accumulated amount of oil decreases enough.

1995. 1993.. 1992. 1998). 2000). the oil film thickness of a particular engine needs to be measured from a firing engine. Scania DSC9 engine with a bore diameter of 115 mm and a stroke of 136 mm was used in the fired engine tests carried out by Mattsson (Mattsson. eddy current sensors. Dearlove and Cheng.. 2000). Josef and Merker. Lubricating oil film thickness under piston rings measured by several authors in test rigs and in motored or fired test engines are tabulated in the study by Grice and co-workers. However. 1998.5 to 5 µm when the engine was cranked
55
. The oil film thickness measurements from a firing engine are very difficult and challenging. six-cylinder. 1993. 1998). The results showed that the oil film thickness between the cylinder liner and a second piston ring. Arcoumanis and co-authors have constructed a reciprocating test rig for lubrication studies with a normal force that varies alongside the applied load. Barrow et al. 1995). (Richardson and Borman. Oil film thickness measuring techniques are presented in the References (Moore. The oil film thickness measurement results show minimum values in the order of 1–2 µm just after the TDC and BDC locations and maximum values in the order of 5–11 µm during the mid-stroke of the piston motion (Arcoumanis et al. show that a thicker oil film can be reached by increasing the engine speed or the oil viscosity. Works by several authors. 1997)... The oil film thickness can be measured by measuring the distance between the piston rings and the cylinder with common contactless sensors like capacitive sensors. 1998. 1995). 1995. in order to understand the correlations between the engine operating conditions. or sensors based on the Hall-principle (Josef and Merker. 1996. at the top reversal position of the ring. Shenghua et al. or by the decrease of the load (cylinder pressure) or temperature. oil film thickness and further the physical phenomenon occurring in the lubricant-surface interaction. Frølund and Schramm. Harigaya et al.. on the thrust side of the piston was from 4.. with high demands and limitations on the tested parameters and on the engine operation conditions. inductive sensors. 1995. (Takiguchi et al.. The degree of the influence of these factors is different and their interaction in a firing engine has an important impact on the lubricating oil film thickness. using test rigs and motored engines. The minimum oil film thickness presented in the table varies from 0 µm to 12 µm and the maximum film thickness between 2. 1997. the position of the piston ring sample during the stroke and the crankshaft speed. Eilts and Wachtmeister. The laser-induced fluorescence technique (LIF) has taken over the oil film thickness measurements (Frølund and Schramm. turbo-charged. 1990).5 µm and 24 µm (Grice et al. A direct-injection.However. results that do not completely follow the trends expected from the theory have been published. Nakayama et al.

5 µm. indirectinjection diesel engine with a bore diameter of 72 mm and a stroke of 72 mm was used as the test engine. • The ring oil film thickness in the piston´s upward strokes was thinner than in the preceding strokes and the oil film thickness tended to become similar among individual rings. The corresponding film thickness at 1 000 rpm and 1 500 rpm and no load was not higher than 0. • The idle running conditions increase the oil film thickness for the top ring during the expansion stroke. In a previous study by Takiguchi and co-workers. naturally aspirated. This was a result of the scarce oil supply to each ring around the BDC. is often reversed. The results showed the following trends of the oil film thickness (Takiguchi et al. and this concerns both the
56
. and between the thrust and anti-thrust sides. • The trends of a computer modelled top ring oil film thickness and the measurements showed good agreement when the speed and load changed. 1995).slowly. The oil film thickness varied markedly from stroke to stroke. It was found that the variation in the oil film thickness was greater than the variation in engine speed or the load. four-stroke. and on the anti-thrust side around the exhaust TDC. In the conclusion of the study the author states among other things that: • The oil film thickness increases almost linearly with speed during the compression stroke. the oil film thickness was measured on both the thrust and anti-thrust sides of the piston at a location of 33 mm from the top reversal point of the ring.. • The piston ring oil film thickness was affected to a higher degree by the amount of lubricating oil supplied than by the engine speed or load. and in the intake stroke on the anti-thrust side. but the calculated values were about 3 to 6 times thicker than the measured values over the speed and load range (Mattsson. A single-cylinder. • The oil film thickness of the piston ring increased markedly in the expansion and intake strokes on the thrust side. This was due to the large amount of lubricating oil supplied from the piston skirt to the oil ring around the compression TDC and exhaust TDC. 1998): • The tendency that the oil film thickness generally increased as the engine speed increased and/or the engine load decreased.

2000. • The effect of the tangential tension of the oil control ring and the width of the ring on the oil film thickness appears mainly around the top dead centre. Additionally. The amount of lubricating oil supply varied markedly with the piston´s slap motion. 2000). Further the authors state that: • A probable cause for the increase in the oil film thickness of the top and second ring on both thrust and anti-thrust sides during the expansion stroke is the restrained access of the gas and thus lower pressure working on the backside of the ring than the gas pressure working on the ring´s sliding surface (Takiguchi et al.. 2000). The oil film thickness increases as the tangential tension is reduced.
57
.. The importance of the oil film thickness measurements in fired engines is not only in the numerical values of the oil film thickness. Seki et al. (Seki et al.. 2000). Later works by Takiguchi and co-workers and Seki and co-authors repeat the above listed trends and emphasise the influence of the piston slap on the ring´s oil film thickness (Takiguchi et al..thrust and anti-thrust sides. which significantly affects the of oil film thickness on the scraper and compression rings. it is fundamental to increase the knowledge about the actual phenomena that occur in the piston assembly and influence the piston ring and cylinder liner contact and thus the piston ring lubrication.

and piston skirt and cylinder liner on the other hand. as the brake thermal efficiency (ηe) of an internal combustion engine can be expressed as
ηe = ηi • ηm = ηt • ηr • ηm
(8.1 Ring friction fundamentals
(8. pf ) can be expressed as the difference between the indicated mean effective pressure (pi ) and the brake mean effective pressure (pe ) of the engine. and wear.8. Expressed in terms of the cylinder pressure of the engine.2)
The sliding contact between a piston ring and a cylinder liner hosts a variety of different friction mechanisms during one working cycle of the engine. Experiments by Takiguchi and co-workers with tworing and three-ring pistons have shown that the number of rings influences the frictional behaviour of the ring pack. As the frictional work losses form a significant proportion of the total mechanical losses.1)
where ηi is the indicated efficiency. the frictional losses can be regarded as highly important for the brake thermal efficiency of an engine. The issue of friction is highly relevant.
58
. The ring friction is determined by the ring load. ηt the cycle efficiency and ηr the relative efficiency (Maass.. Owing to the variations in load.1. 1996). the frictional losses (friction pressure. 1979). The ring load comprises the ring pre-tension and the gas forces acting on the back-side of the ring. speed and counter surface effects. the lubrication conditions in a ring/liner contact are strongly transient. the surface properties and the lubrication conditions as determined by the sliding velocity and the viscosity and availability of the oil. behaviour. Friction of piston rings and skirt against cylinder liner
This section covers various aspects of the presence of a coefficient of friction between piston ring and cylinder liner on the one hand.pf = pe 8. which is reflected by variations in the friction. ηm the mechanical efficiency. as follows:
pi . but that the total tension of the piston rings in the ring pack finally determines the friction losses (Takiguchi et al.1 Friction of piston rings against cylinder liner
8.

The frictional behaviour of the piston. speed. in which the (i) coefficient of friction. 1998). or (iv) the angular position of the crankshaft. will determine the precise friction mechanism at the dead centres and the effect of the elastohydrodynamic lubrication between the dead centres of the piston motion. Coy. In addition to the oil supply. and the geometry of the sliding contact. 2001. which is strongly affected by the hydrodynamic lubrication conditions between the TDC and BDC locations.. 1995. 1998. has been found to increase with an increase in the viscosity of the lubricating oil (Hamatake et al.... the momentary friction mechanism depends on the load. The availability of oil. by the lubrication effect of the graphite phase and by the oil reservoir provided by the graphite phase of the material (Glaeser. 1997.. Under conditions of lubricating oil starvation. Recent studies by several authors have shown that. friction modifiers like molybdenum dialkylthiocarbamate (MoDTC) and Ca-additives that may form wear-resistant layers of CaCO3 strongly control the boundary friction conditions.In terms of the commonly used Stribeck diagram. in particular. The maximum friction force. Friction curves of any of the above-mentioned types
59
. has been found to decrease with increasing oil viscosity. which occurs under conditions of mixed lubrication in the vicinity of the TDC. in terms of either poorly lubricated or fully flooded conditions... Hamatake et al. Saini et al. As a short summary of the friction mechanisms active in a ring/liner contact. 1998). Korcek et al. the lubrication conditions in a ring/liner contact experiences strong and rapid movements on the horizontal axis of the diagram (Taylor. 2001). piston ring and cylinder liner can be expressed in several different ways. Anti-wear additives like ZDDP. Durga et al. grey cast iron provides certain reduction in the friction forces. 1998. the organomolybdenum compounds MoDTC and molybdenum dithiophosphate (MoDTP) strongly reduce the coefficient of friction under boundary lubricated sliding conditions (Glidewell and Korcek. or (ii) the friction force is plotted with respect to either (iii) the piston stroke. 2001). actual lubricant viscosity. 2000. 2001). The oil formulation strongly affects the formation of boundary layers on the mating surfaces.. the friction mechanism active in the vicinity of the dead centres of the piston movement is a combination of boundary or mixed lubrication with an additional lubricant film squeeze effect at the dead centres. while the friction mechanism active in the mid-stroke of the piston motion is hydrodynamic lubrication (Wakuri et al. The most detailed analyses comprise friction curves. 1992). Tung and Tseregounis. Work by Zhang and co-workers has shown that organomolybdenum-sulphur compounds decompose to form lubricious MoS2 on metal surfaces at high temperatures (Zhang et al. Arcoumanis et al. while the friction pressure pf. 2001. depending on the purpose of the tribological analysis.

when minimising the friction losses.. The oil film thickness is. The interaction between the ring and liner wall becomes more and more important as the oil film thickness decreases. Various models for shear thinning behaviour have been developed. for the midstroke region of the piston movement can be useful for the assessment of variations in the lubrication conditions following from different tribological parameter combinations. but it is not able to completely separate them. or (vi) the friction force. Weighted average friction force measurement results can furthermore be expressed as (ix) the friction pressure.
8. Nevertheless. The phenomenon is called shear thinning and is usually taken into account in computer simulation. and the probability for wear. The simulation model works as a hydrodynamic model
60
. Surface characteristics of piston ring face and liner wall surfaces are of great importance for the conditions of mixed lubrication.serve for tribological analyses of the lubrication conditions. For enabling calculations of the frictional power loss on the basis of a friction coefficient curve for a piston/liner pair.2 Friction simulation
A strict relationship prevails between oil film thickness and friction. in terms of (vii) the coefficient of friction or (viii) the friction force.1. not all computer models include friction calculations. thus small friction loss levels require compromises between hydrodynamic and sliding contact friction. in the ring/liner or piston/liner contact. decreased in order to make the hydrodynamic friction loss as small as possible. An oil film exists between the two interacting surfaces. Weighted average friction measurement results. the normal force between the piston assembly and the cylinder liner needs to be known in detail. It is evident that surface contact will occur at thin oil films. for the entire working cycle contain less information for use in lubrication analyses but they are more useful for the assessment of the frictional power losses of the engine. Wakuri and coworkers note that since perfectly hydrodynamic lubrication in the ring pack cannot be ensured. 1995). The oil viscosity undergoes decreases when the temperature increases. in terms of (v) the coefficient of friction. Surface contact in turn increases the total friction losses. where surface asperity contact occurs but is not dominant. and the viscosity decreases when the shear stress increases.2) above). pf. The friction models that have been included in the computer simulation vary essentially depending on what the author of the model considers as significant. which is equal to the difference between the indicated mean effective pressure and the brake mean effective pressure (see formula (8. Weighted average friction measurement results. Computer simulation of ring performance generally includes mixed lubrication models. the theoretical estimation of the friction should always include a mixed lubrication model with asperity contacts (Wakuri et al.

The combustion gas acts on the back of the ring and presses it against the cylinder liner.. 1996). The computational models used are an accurate model that takes into account the inertia and squeeze film effects. which also reduces the friction near the dead centres.until the oil film thickness decreases to a certain value. where the total engine friction is investigated. The influence of laser-textured cylinder bore and piston ring surfaces on the friction has been studied by Ronen and co-workers. Studies of two-ring pistons in spark-ignited engines have been made by Takiguchi and co-workers in order to clarify their characteristic oil film thickness. the oil ring friction seems to be most sensitive to surface roughness variations (Sui and Ariga. Furthermore. after which the mixed lubrication model is included in the simulation. Sui and Ariga have investigated the effects of the ring surface topography on the ring/liner interface friction. The authors found out that though two-ring pistons have greater blow-by than three-ring pistons. compared to three-ring pistons. and an approximate model that ignores these two factors.. the gas pressure on the back-side of the other piston ring is so low. Owing to the sealing capability of the first compression ring. It is shown that instantaneous clearance and friction results obtained from the approximate model may differ considerably (Ronen et al. According to the results. the authors conclude that the piston ring force can be reduced regardless of the number of piston rings. The results include simulations for fully warmed-up conditions and cold-start conditions. When using a computer model where only the inertia effects are ignored. 1993). that the frictional power is not affected by load variations (Reipert and Voigt. The sliding frictional power at the top ring surface is increased at increasing load. 2001). Furthermore. A model for the prediction of the engine friction has been presented by Taylor. The friction reduction results in an increase in the oil film thickness. which is based on the mixed lubrication concept. The top ring experiences a higher frictional power owing to surface contact rather than hydrodynamic action. the total engine
61
. 2001). The lubrication model was extended for studies on the ring-pack friction in firing engine conditions. They developed a ring-pack friction model. The simulation results were verified in a moving liner test bench. there is no saving in computer time. which leads to a very small increase in oil consumption. the oil film thickness for two-ring pistons in turn becomes thinner. by reducing the tension of the piston rings (Takiguchi et al. This critical value of the oil film is determined from the roughness of the interacting surfaces. ring friction and oil consumption. Results obtained from these tests indicate that up to 9 % reduction in the friction loss is possible by changing the surface pattern.

the friction graphs comprise characteristic peaks at the start and stop locations.. Dearlove and Cheng have measured the coefficient of friction of a chromium plated piston ring with a barrel profile oscillating against a polished and honed cast iron cylinder liner sample under lubrication with five different oils at 30°C temperature. Examples of different ways in which the friction measurement results can be expressed are mentioned in the last paragraph of Section 8. and they report average coefficients of friction of approximately µ = 0.07 at middownstroke. while other authors present friction graphs that lack any peaks of significance at the start and stop locations. This difference may more strongly reflect differences in the response of the friction force measurement arrangements than in the tribological phenomena studied.1.1. Experimental investigations by different authors on the tribological properties of the piston ring / cylinder liner have employed a piston ring or a ring segment sliding under reciprocation against a part of a cylinder liner or a complete floating cylinder liner. Miniaturised or downscaled tests for the evaluation of the tribological properties of the ring/liner/lubricant system are less expensive than the engine tests. at 200 rpm engine speed and 40.
8. The friction force has been measured either from the ring or from the liner. the position of the piston ring sample during the stroke and the crankshaft speed.3 Measured friction forces and coefficients of friction
Piston rings are often . 1993. (Patterson et al.80 N normal force. A crank or eccentric mechanism. The friction curves measured with the installation show peak values just after the TDC and BDC positions of the piston ring. 1995). At an engine speed of
62
.friction immediately after a cold-start is four to five times higher than at fully warmedup conditions (Taylor. or an electromagnetic actuator has been employed for achieving a horizontal or vertical reciprocating motion of a piston ring or cylinder liner specimen in the test rigs..and must often be ..tested in motored or fired engine bench tests. 2000).1. 1997).. while more advanced studies have employed more application-oriented test equipment or motored or fired test engines. Noorman et al. Arcoumanis and co-authors have constructed a reciprocating test rig for lubrication studies with a normal force that varies alongside the applied load.. higher friction forces with higher velocities and higher friction forces but lower coefficients of friction with higher loads (Arcoumanis et al. In some works. The simplest test set-ups are based on commercially available oscillating wear test rigs (DIN 51 834) and pin-on-disc test equipment for general tribological test purposes. and different phenomena can be investigated more easily without interactions from other phenomena acting in parallel.

400 rpm.03. The authors present indications of mixed and hydrodynamic lubrication. The results by Glidewell and Korcek clearly show the necessity to consider the complete formulation of a lubricating oil when analysing the results of piston ring friction measurements (Glidewell and Korcek. 1993). 1995). an observation which is in agreement with work by Hamatake and co-workers. 1998). Due to a lower oil viscosity. 1993).0.05 were obtained with oils containing a molybdenum dialkylthiocarbamate (MoDTC) friction modifier.125 at 202° temperature.
63
. weighted average coefficients of friction in the range of µ = 0. the average coefficient of friction was approximately µ = 0. The normal load applied on the tests was quite low. 1999a. 2001).095 at 100°C temperature and approximately µ = 0. lubricated with a 1 % solution of stearic acid in white oil.11 within the same temperature interval. who present lower coefficients of friction as associated with oils of lower viscosity (Hamatake et al. while coefficients of friction of µ = 0. a lower coefficient of friction was recorded by Sui and Ariga at 93°C temperature than at 30°C temperature (Sui and Ariga. has been presented by Sui and Ariga. In tests with pieces of chromium-plated cast-iron piston rings sliding in an oscillating mode against samples of honed cast-iron cylinder liners.. In the measurements. while sliding at the latter temperature gave rise to scuffing. and at 600 rpm the value was µ = 0. Glidewell and Korcek have presented results obtained in reciprocating tests with a molybdenum-coated cast-iron piston ring sliding against a cast-iron cylinder liner sample.10. the corresponding coefficient of friction was approximately µ = 0. Galligan and co-workers have measured average coefficients of friction of approximately µ = 0. 1999b).. while the above results presented by Sui and Ariga seem to represent lubrication conditions corresponding to the right from the minimum value of the same curve. A minor although noticeable increase in the friction coefficient at mid-stroke. When the oil was replaced by a fully formulated 15W50 motor oil. who have compared the friction force of a ring pack at 650 and 1 500 rpm (Sui and Ariga.. The above results presented by Dearlove and Cheng seem to represent lubrication conditions corresponding to the left from the minimum value of the Stribeck curve. owing to increased hydrodynamic shear loss with increasing speed.. which may explain the strong reduction in the coefficient of friction when increasing the engine speed from 200 to 600 rpm (Dearlove and Cheng.0. which were carried out with different oils at 100°C temperature.11 were obtained with oils not containing a friction modifier. while sliding at approximately 248°C gave rise to scuffing (Galligan et al.04.06...

1998). Tests with piston ring and cylinder liner samples performed by Tung and Tseregounis using a reciprocating rig gave average coefficients of friction of approximately µ = 0.. surface quality and surface material (Durga et al.. the authors have not very much focussed on the analysis of different lubrication regimes within the Ushaped friction graphs (Akalin and Newaz. have been presented by Akalin and Newaz.. Dearlove and Cheng.10. 1998).. Dearlove and Cheng. and lower values down to µ = 0.15 for the TDC and BDC positions.0. For a chromium-plated ring against a cast-iron cylinder liner lubricated with SAE 5W30 motor oil at room temperature.10.03 in the midstroke regions (Akalin and Newaz. the respective values depending on the actual lubricant. Andersson..
Fig. for 10 Hz oscillating frequency of the piston ring (after Andersson. 1998).08 with low additive oils. and values down to µ = 0. 1998.. 8. Durga et al.0.03 for oils with high Mo concentrations and high Mo/S or Mo/Zn ratios (Tung and Tseregounis.1. 2000).15 and mid-stroke values in the range of µ = 0. Akalin and Newaz show peak values of µ = 0..02.0....12. However. The curve corresponds to 360 degrees of crankshaft rotation.0. 2002.05. Example of a U∩-shaped friction graph. or U-shaped (mostly U∩ shaped) friction graphs (Fig.Friction graphs from tests with pieces of piston rings sliding in an oscillating mode against samples of cylinder liners that comprise clearly higher coefficients of friction in the vicinity of the TDC and BDC locations. and Durga and co-workers. 8.
64
. 2002). Durga and co-workers have reported peak values at the TDC and BDC locations in the range of µ = 0. Glidewell and Korcek. Andersson. Glidewell and Korcek. 1995.1). 1998.

1. on the effect of exhaust gas re-circulation (EGR) has shown an increase in the wear of the piston rings and the coefficient of friction when using EGR. which arises from the presence of an excess of combustion chamber deposits in the ring groove was found to be more severe with rings with a high angle of torsion and low actual side clearance (Akimoto et al. The half-sticking. 2001). 1998).. The phenomenon of half-sticking partly locks the top ring in the groove for a short period of time.. 1993). The increase in friction and wear is assumed to occur from excessive carbon deposits in the ring groove following the unfiltered exhaust gas re-circulation (Urabe et al.
8..10. especially without soot filtering (Urabe et al. are supported by earlier analyses performed by Sui and Ariga (Sui and Ariga. on ceramic coatings in a reciprocating test rig with a flat-on-flat geometry have show that a chromia (Cr2O3) coating applied on both sliding surfaces and lubricated with a polyalphaolefine oil can give a coefficient of
65
. the surface finish is of great importance for the lubrication conditions and the frictional behaviour. For decades now.. Ma and co-workers have shown that the coefficient of friction is strongly reduced during the running-in stage of the engine operation (Ma et al. Hamatake et al..10 and weighted average values of µ = 0. and Hamatake and co-workers. both for suppressing wear and for reducing friction.15. Hard chromium coatings with oil-retaining porosity or channels on piston rings are good examples of a traditional beneficial piston ring coating. and Priest and Taylor have shown the same effect in terms of fuel consumption (Priest and Taylor. who have investigated a phenomenon of halfsticking of the top piston ring by using a visual access device in the shape of a sapphire glass window in the cylinder liner.A conclusion of the above piston ring friction measurements is that the friction curve obviously has a U-shape. The influence of the surface finish of the piston assembly on the total frictional losses of an engine has been studied by engine tests by Wakuri and co-authors. who have shown that the frictional losses increase when run-in pistons. mid-stroke values in the order of µ = 0.0. Work by Urabe and co-workers.11.. piston rings have been coated. Investigations by Glaeser and Gaydos.02. The benefits of reducing the surface roughness.0..4 Effect of piston ring surface finishing and coating
As for any tribological surface. 1998). 2000). 2000)... in particular for reducing the friction and wear under mixed lubrication conditions. with peak values in the range of µ = 0.. Another approach for increasing the understanding of piston ring friction has been presented by Akimoto and co-authors. piston rings and cylinder liners are replaced by new components that need to be run-in (Wakuri et al..04. 1998. 1995)..0.

which considers the aspects of both friction reduction and scuffing suppression.. Their study was related to the development of the adiabatic diesel engine (Glaeser and Gaydos. the coefficient of friction in the beginning of an oscillating test is lower (µ = 0. 1999a).. Rvk (reduced trough depth) and Ra (arithmetic average) have been used as measures on the oil-retaining capability of the groove pattern that has been produced onto the liner surface by honing. As described below. and report mid-stroke values in the order of µ = 0.. which consists of a fairly flat base surface with a network of deep scars. after a certain sliding distance the coefficient of friction is at the same level irrespective of the difference in initial surface quality (Galligan et al.1.07 µm with a 5W30 oil and engine speeds of 100 and 450 rpm. Durga and co-workers have investigated the effect of different surface roughness values of a cast iron cylinder liner on the coefficient of friction.3 µm). The oil-retaining volume of the honed cylinder liner surface is of substantial relevance for the tribological performance of the system. surface roughness parameters like the RSK (profile skewness). Coatings on cylinder liner surfaces may offer further benefits. the surface quality of the cylinder liner largely determines the scuffing resistance of the cylinder and piston assembly combination.05.3 µm and µ = 0. has been found appropriate for the lubrication of the piston assembly..1 against 0. 1998).5 Effect of cylinder liner surface finishing and coating
A plateau-honed cylinder liner surface profile.13) with a highly polished cylinder liner than with a liner with a standard surface finish.08 with an Ra value of 0.08 at 260°C temperature.friction in the range of µ = 0. from which point of view too smooth a surface liner may be unfavourable.0.12 with an Ra value of 0. gave lower coefficients of friction than the honed cast-iron cylinder liners under identical test conditions (Durga et al. For some decades. 1993). From the above investigations on the influence of the surface quality of the cylinder liner on the engine friction it can be concluded that an optimum surface quality can be established. in the sections concerning scuffing. even with a fairly coarse surface finish (Ra 0. According to the work by Galligan and co-authors. However.
8. Plasma-sprayed coatings.
66
.

1976) are likely to be pronounced on areas subjected to higher contact pressure. the frictional losses of the engine and the friction force variations in actual engines. Friction force peaks recorded at the TDC and BDC locations of the piston motion were about twice as high as the uniform friction force at mid-stroke at an engine speed of 100 rpm. At 500 rpm engine speed the force peaks were not visible. 1989. 700 and 1000 rpm engine speed. 420 and 380 N in firing engine tests with SAE 30 oil at 70°C temperature at 500. 1995). probably owing to the mass inertia of the floating liner (Clarke et al. by using the floating cylinder liner or movable bore technique. the piston skirt design. The friction force readings are..
8.8. In the case of severe roundness errors. A
67
. Friction measurements by different authors have been carried out with piston rings in firing engines. however. highly useful. Wakuri and co-workers report friction force measurement results with peak values of 570. the piston assembly friction has been studied by motored engine tests. a concise friction coefficient curve cannot normally be established. the piston design and the tilting action of the piston. The corresponding coefficient of friction was µ = 0. With a SAE50 oil the friction force peak value was 300 N and with a SAE10 oil 560 N at 70°C temperature and 700 rpm engine speed.6 Effect of cylinder liner out-of-roundness
Deviations from cylindricity of cylinder liners cause local variations in the contact pressure between the piston rings and the cylinder. As the corresponding normal force on a particular piston ring is known only to certain degree. particularly with stiff piston rings and high crankshaft speeds. areas of particularly low contact pressure between ring and liner are subjected to increased risk of combustion gas blow-by. the cylinder liner is allowed to move axially for the minute distance that is necessary for enabling force measurements from the liner. Alternatively to employing engine tests using the floating liner technique. In the measurement arrangements. as they express the axial loads on the rings. The wear of the cylinder liner and the likelihood of bore polishing and piston ring scuffing (Munro. Clarke. Sherrington and Smith have reported on the development of a floating liner technique and friction graphs recorded using their equipment with loading by compressed air.1.2 Friction of piston skirt and piston rings against the cylinder liner
The friction between the piston skirt and the cylinder liner is controlled by the diameter clearance. the surface roughness and the preconditions for lubrication (Röhrle. 1990).08 at the intermediate engine speed.

Durga and co-workers have performed motored engine tests... 1998). 1998).. The maximum friction force was found to decrease with increasing oil viscosity. and the temperature of the SAE 30 oil used was 70°C. 1998). 2002). the friction pressure pf was about 40 kPa at 0.higher gas pressure peak (7 MPa) gave a higher friction force peak value (560 N) than did a lower (5 MPa) gas pressure peak value (310 N).48 µm) in motoring torque tests up to 6 000 rpm. although with lower contact pressures. but not above this speed (Durga et al.. when the cylinder pressure was about 6 MPa.07 µm) cast-iron cylinder liner surface gave rise to lower frictional losses than a rougher liner surface (Ra 0. while the friction pressure pf was found to increase with an increase in the viscosity of the lubricating oil (Hamatake et al. the crankshaft speed 1 000 rpm. that a higher velocity improves the conditions for hydrodynamic lubrication. and that a higher cylinder pressure gives higher frictional losses in the piston and cylinder contact. These findings by Wakuri and coworkers indicate that the oil viscosity supports the formation of the oil film between piston and cylinder.25 % of full load. At the same engine speed. 2001).
68
. at the piston skirt. and they report friction peak forces of 400 N immediately after the moment of combustion.. In a review on lubrication models for engines.. and from the response of the two floating liner systems to the oscillating frequencies applied. Hamatake and co-workers have studied the friction force of the piston (∅ 105 mm) of a diesel engine with a floating liner. 1995). Work by Golloch and co-authors on a ∅ 128 mm diesel piston assembly has shown maximum friction forces up to 2400 N for a mean cylinder pressure of 20 bar. in which cylinder liner coatings were found favourable. particularly in the ring/liner contact (Wakuri et al. A smoother (Ra 0. 1 000 rpm engine speed and one-quarter full load (Coy. Coy presents peak friction forces in the order of 450 N and the corresponding friction mean effective pressures in the order of 42 kPa from engine tests with 10W50 oils at 70°C temperature. a maximum cylinder pressure of 170 bar and a crankshaft speed of 800 rpm (Golloch et al.
8. hence more favourable lubrication conditions. The difference in the above findings may arise from different cylinder pressures and cylinder diameters.2. Similar work by Urabe and co-workers on a ∅ 108 mm diesel piston assembly has shown maximum friction forces up to 320 N for a maximum cylinder pressure of 60 bar and a crankshaft speed of 1 200 rpm (Urabe et al..1 Normal operation conditions
Under normal operating conditions the lubrication conditions for the piston skirt and cylinder liner surfaces resemble the lubrication conditions of the ring pack.

1976) than today. The scuffing phenomenon was a larger problem a few decades ago (Aue.e. On the piston ring and cylinder liner surfaces evidence of scuffing may be found in the shape of wear scars indicating. may contribute to the scuffing. by wiping off the oil from certain locations of the cylinder liner.and three-body abrasive wear and plastic material flow at the ring and liner surfaces (Galligan et al. which consists of local micro welding. 1999a. while a good availability of lubricating oil increases the time before scuffing occurs (Galligan et al. The local temperature increase can. with the purpose of scraping off excessive carbon deposits from the piston before the
69
. 1999b). Experimental investigations by Galligan and co-workers have shown that an increase in operating temperature.2. abrasive ploughing and the adhesive transfer of work hardened cast iron to a chromium-plated piston ring. The cylinder liner bore polishing can be a consequence of two. and a "white layer" that indicates that the temperature has locally exceeded 750°C (Lacey and Stockwell. in turn. or "hot spots". above the top piston ring. and the presence of martensitic transformation on the cylinder liner surface (Shuster et al. 1999).2 Scuffing
Piston ring scuffing is a randomly occurring phenomenon. 1999). consequently. plastic deformation. 1999b). and the occurrence of flash temperatures. between a piston ring and a cylinder liner.. or material adhesion. which instantly leads to an increase in the contact pressure in the ring/liner contact. Excessive deposits of carbon on the top land of the piston. there is risk for an increase in scuffing problems when the oil film thickness in the ring/liner contact is reduced for environmental protection reasons (Aue.. 1976).8. Lacey and Stockwell. 1999). One of several locations. i. in the ring/liner contact. Metallurgical investigations by Shuster and co-workers on initial scuffing failures have shown the presence of minor iron-based particles on the face surfaces of Mo. e. 1999a. cause the formation of local thermal expansion of the cylinder wall material towards the bore ("thermal bump")... an increase in load and an increase in the oscillating frequency of the test samples shorten the time before scuffing occurs.g. However.and Cr-coated piston rings. This problem may be overcome by the use of a steel ring of slightly smaller diameter than the cylinder bore diameter at the top end of the cylinder liner. 1976. The lubrication conditions may be poor either owing to a large surface roughness (low λ value) or to cylinder bore polishing (lowered Rvk value) that has erased the honing marks and. the oil reservoirs from the surface. where the risk of scuffing is large is between TDC and mid-stroke where the product F×v reaches a maximum (Willn and Brett. The scuffing phenomenon is normally preceded by conditions of locally starved lubrication.

1976).2. For similar reasons as for piston rings. 1999). piston seizure can be the result of insufficient piston/liner clearance.deposits form a hard layer (Amoser. In practice. Wang and Tung have presented the results of a scuffing resistance study on various candidate coatings for aluminium piston skirts in aluminium cylinder liners (Wang and Tung. while at a coefficient of friction of µ = 0. From this point of view. piston skirt scuffing may become a problem. iron plating of the piston skirt has been the prime solution against piston scuffing. 1998). where large differences in time-to-scuff under otherwise identical sliding conditions have been shown (Galligan et al.0.. and can be preceded by piston scuffing. and particularly in the case of new engines. particularly with aluminium pistons in aluminium cylinder liners. Recent work on the detection of piston ring scuffing has involved the use of acoustic emission detectors. and thus controls the availability of the oil in the ring/liner contact (Lacey and Stockwell. However. 2000). Carbon deposits in the piston ring grooves and the ring sticking owing to the deposits increase the probability of scuffing (Munro. above which the reaction layer breaks down and the probability of scuffing increases dramatically. Recently.11 or higher. Alternatively.. The seizure comprises jamming of the piston in the cylinder bore owing to strong adhesion between the mating materials. which controls the evaporation of the oil or part of its constitutes at high temperature. Additional protection against scuffing is achieved by engine oil additives. for each reaction layer system formed from surfacereactive additives there exists a maximum temperature. it can be concluded that piston ring scuffing is likely to occur if the surface roughness is too high or too low.10 scuffing did not occur (Durga et al. The effect of the reaction layers in the ring/liner contact is obvious in recent studies with different oil formulations. Poor running-in properties of piston rings. which promote the formation of an effective boundary lubrication film and thus suppress the scuffing phenomenon. Durga and co-workers have reported from LS-9 Scuff Resistance Tests that scuffing occured at a coefficient of friction of µ = 0. by which the scuffing damage is subdivided into scuffing origin. piston seizure is the result of tribological and/or thermal overloading of the piston ring pack and skirt. 1999).
70
. owing to surface layers that are too wear resistant. 1999a. irreversible scuffing and severe scuffing (Shuster et al. 2001)..09. Since the introduction of aluminium cylinders in automobile engines three decades ago.
8..3 Piston seizure
Piston seizure is a phenomenon that has not attained much attraction in terms of research. may increase the risk of scuffing. 1999b).. Another feature of particular interest is the volatility of the oil at the cylinder wall.

Wear in the piston-ring-liner system
9.5) and wear inversely proportional to the oil film thickness (Coy. ring and cylinder surfaces. 1998). Concerning most tribological applications. leading to piston ring scuffing that comprises high friction forces and the formation of severe wear scars on the piston.. the wear can be estimated from changes in relevant surface roughness parameters representing certain proportions of the piston ring face surface area (Sherrington and Mercer. literature on the influence of the tribochemical wear on the overall wear of piston rings is only available to a rather limited extent.. The wear of piston rings and cylinder liners can be accelerated by three-body abrasive wear caused by minor abrasive particles in the lubricating oil. the overall wear rate can be tribochemically accelerated by aggressive components in the lubricant that have been entrapped in the ring zone. Under conditions of poor lubrication. 2000). In addition to the two-body and three-body abrasive wear. while in reality the wear process is significantly more complicated (Gupta. while sliding under less favourable conditions in the vicinity of the dead centres of the piston motion cause mixed lubrication (λ = 1. 1999). Archard and Hirst. Aggressive combustion products are formed in particular when highly sulphuric fuels are used. 1998). For low wear rates. Experiences of chromium plated piston rings show that they offer good protection against wear caused by acidic combustion products (Federal-Mogul. For high wear rates. expressed by the formulae presented by Archard. Preston. surface degradation of piston rings can take place due to blow-by of hot gases from the combustion chamber. Alternatively.1 Wear of piston rings
It is commonly assumed that the wear of piston rings proceeds according to a mild mechanism of mild two-body abrasive wear against the cylinder liner. the wear volume can be determined from macro geometrical changes or mass loss. the wear volume of piston rings can be determined by comparison of surface roughness profiles or cross section profiles before and after the tests (Shuster et al. 2001). Kauzlarich and Williams. strong adhesive forces between the piston rings and cylinder liner may occur. 2001. The contaminant particles causing the three-body abrasive wear can originate from the oil sump or from the combustion chamber. Rabinowitcz or Holm. where the temperature of the
71
.9. In addition to sliding wear. conditions of hydrodynamic lubrication at the mid-stroke region of the piston motion give rise to full film lubrication (λ > 5) and zero wear. As presented by Coy in his qualitative wear transition model..

1 Running-in wear of piston rings
The most intensive wear of the piston ring pack and the cylinder liner normally occurs in the running-in.
9. or break-in.combustion gas is in the excess of 2 000°C.2. and that approximately 84 % of the ring wear occurred during the first approximately 22 minutes of operation (Perrin et al.
9. and the dynamic axial loads on the rings. subsequently to the running-in stage of the piston rings and cylinder liner. In engines where ring deterioration owing to blow-by is likely to occur. This phenomenon is further described under Section 9. the use of molybdenum or similar heat-resistant coatings is essential (Brauers and Neuhäuser. the wear rate of a cylinder liner is approximately 12 times higher during the first hour of operation than during the subsequent two hours (Henein et al. 1991.1. Ma et al. and the wear of the cylinder liner decreases when the piston rings become smoother by runningin. under favourable conditions this self-stabilising process leads to a decrease in the wear rate of both the piston rings and the cylinder liner (Hu et al. The blow-by can cause local melting or hot gas erosion damages. According to experimental work presented by Henein and co-workers. 1998). Unless thermomechanical fatigue fracturing of the piston rings or their surface layers occurs. 1998).. the low level of the wear rates can continue until the components are taken out of use. 1989). in conjunction with
72
.2 Steady-state wear of piston rings
In steady-state conditions.1. stage of the engine. wear occurs in the ring/groove contact. on the rings. the wear rates of the respective counter bodies is a fraction of the level experienced during the running-in. The abrasive wear of the piston rings decreases when the cylinder liner surface irregularities become smaller. or burn scars. during which the most predominant surface profile peaks are worn off by the counter surface.. and the surfaces eventually obtain improved conformity (Priest and Taylor.1. Consequently.
9. Wear tests with neutron bombarded compression rings in fired engine tests including a gamma ray spectrometer have shown that the piston ring wear rate during the start-up period was up to 45 times the steady-state wear rate. 2000)... 1995).3 Wear of ring flank surfaces
Owing to the relative motion between the ring flank surfaces and the grooves in the pistons.

Wear of the piston rings and the cylinder liner is perhaps the most difficult phenomenon to implement in a calculation model. They have modelled the top ring of a Caterpillar 1Y73 engine and compared the results over 120 hours of engine running. The oil rings work under an oil film thickness that is thinner than that of the compression rings on average. Ring twist is also an important
73
.1)
Where k is the wear factor. The wear factor. 2002). They point out that piston ring designs with emphasis on wear resistance may be non-optimal considering the lubricational and frictional properties (Priest and Taylor. xs the sliding distance and V the worn volume.1. the oil and operating conditions. and for this reason the wear of the oil rings cannot be neglected. Therefore. W the normal load.4 Wear of oil rings
The tribology of the oil rings is not analysed separately in greater detail in this work. The wear factor has to be determined by empirical methods. from bench tests. wear comprises less understood phenomena than friction or lubrication.details on the wear of the ring grooves. Hard carbon particles from the combustion process have been observed to cause wear on the spring expanders of three-piece oil rings. Priest and Taylor have investigated piston ring wear modelling. 1998). wear is a phenomenon to be included in a realistic model. The time required for simulation increases when wear models are included in the simulation software.
9.1. Some secondary wear is known to occur as the result of interactions between the different parts of compound oil rings (Federal-Mogul. The ring flank wear can be reduced by the use of coatings on the ring flank surfaces. Even though wear might be considered a minor factor in calculation models. it should be remembered that the wear of a piston ring alters the ring profile.5 Simulation of wear of piston rings
The wear rate can be mathematically defined using the Archard wear equation: V = kWx s (9. mainly owing to the lack of specific information available in the literature. Priest and Taylor have examined the correlation between predicted and measured wear results. 2000). In terms of the simulation. Surface nitriding provides a means of preventing the wear of the oil rings (Esser. k is in turn a function of the surface properties.
9. for example. owing to the high spring forces that continuously act on the oil ring. The results suggest that the wear was not equal at all locations along the circumference of the ring. Wear parameters most certainly require empirical data.

Commercial simulation software is limited in terms of the simulation of the wear rate. occurs mainly in the top ring groove. Sui and co-workers suggest that the best way to maintain the ring profile is to consider the surface topography effects at an early stage of the ringpack design (Sui and Ariga..factor to take into consideration in wear modelling. the primary reason being an extensive processing time and challenging implementation of the calculation results. 1993). 2000). This. Ring profile change owing to surface wear is ignored. According to this model. The reasons for the radial motion of the ring
74
. and the wear process is accelerated by poor lubrication and a high temperature. Coy has investigated the wear of the top ring related to the distance from the TDC. 1998).
9. Wear is minimal at full hydrodynamic or elastohydrodynamic (EHD) lubrication conditions. 1998). as has been shown by Coy (Coy. and the authors conclude that especially the top ring profile wear is considerable during the 120-hour test. it is not necessary to consider surface contacts in the simulation. At the hydrodynamic or elastohydrodynamic lubrication regimes. The higher wear rate areas correspond to the areas where mixed and possibly even boundary lubrication occurs (Coy. as they require a wear coefficient that reflects the wear of both mating surfaces. however. Piston ring profile wear has been investigated by Priest and co-workers. in turn. where there is asperity contact. With a transition wear model. the wear model becomes more important. The main reason for the wear is the combined effect of gas forces and radial motion of the ring. The lubrication model used did not determine the cyclic variations in the torsional twist angle of the ring. This feature is. a ring twist pattern was assumed according to a predicted twist angle. commonly called ring-groove wear. The results show good agreement between the predicted and experimental results (Priest et al. If the lubrication conditions change into mixed lubrication. needed for the wear model. the wear rate increases again. Therefore. the surface roughness of the rings and cylinder wall is significantly reduced during the test (Priest and Taylor. Owing to the wear. When the piston approaches the BDC. as nearly no asperity contact occurs in these conditions. Consequently. 1999). Piston ring profile wear and lubricant degradation during a 120-hour cycle was included in the wear model. the ring profile has to be maintained in order to keep the ring friction at a desired level. the wear rate is at its maximum at the TCD and decreases as the piston moves downwards. leads to variations in the ring friction.2 Wear in the piston ring groove
Wear of the parallel surfaces in piston ring grooves.

. the piston skirt and the cylinder liner must fulfil certain requirements of surface smoothness. Wear of the piston skirt normally remains low. and the upper surface becomes rough (Röhrle. Affenzeller and Gläser. 1995. 1995). 1998. Mass forces. Alternatively to wear.1 Mild wear of the piston skirt
As part of the running-in of the piston assembly. mild wear of the thrust and anti-thrust sides of the piston skirt are a normal consequence. 1995. which accelerates the ring groove wear at the ring-groove contact areas. Soft coatings on the piston skirt can be applied for supporting a mild wear process during the running-in of the piston/cylinder pair.. as a consequence of overheating of the top ring or partial seizure of the piston owing to poor lubrication (Willcock. friction forces.3. In continuous operating conditions. and the side clearance between ring and groove increases.3 Wear of the piston skirt
9. the secondary movement of the piston and piston tilt allowed by the piston/cylinder clearance. the ring-groove tribosystem may suffer from ring welding.are the cylinder distortion. ring groove wear in aluminium pistons is prevented by a cast-in ring carrier made of austenitic Niresist cast steel or similar wear-resistant material (Mollenhauer. the lower surface of the ring groove becomes rough and rounded towards its edge. axial ring movement and ring rotation increase the ring groove wear. Röhrle. Federal Mogul. 1996). the width of the groove increases. 1996. the wear of the piston skirt is normally insignificant.3 µm are typical of the piston skirt (Röhrle. In certain diesel engines for passenger cars and all advanced diesel engines for commercial vehicles. For improving the break-in properties of the skirt. (Röhrle. Instationary gas pressure and gas blow-by may cause radial vibrations in the ring. 1997. despite the hardness reduction of the piston material at the operating temperature. and for providing conditions of hydrodynamic lubrication and low wear. As the result of the wear. 1997). as the hydrodynamic lubrication conditions are disturbed only by the reversal of the motion at the TDC and BDC locations. Affenzeller and Gläser.5. 1996). Surface roughness values (parameter not indicated) in the order of 1. Mollenhauer. 1995).
75
. In addition to the deformation of the groove arising from wear.
9.

the risk of lubrication starvation and scuffing is obvious (Dong et al. Wear of cylinder liners occurs as well in the mid-stroke region of the piston ring motion. The erosive loads comprise the mechanical effect of the flushing by hot gases along the upper parts of the cylinder liner surface. Cylinder bore polishing.
9. and to a less extent with the bottom reversal points of the piston rings. Carbon deposits above the ring pack on the piston may significantly increase the cylinder liner wear in the TDC region (Affenzeller and Gläser. which can be subdivided into light. chemical. particular at low cylinder surface temperatures. 1995. A high sulphur content of the fuel can increase the proportion of tribochemical wear of the cylinder liner dramatically.4 Wear of the cylinder liner
Wear of the cylinder liner is caused to a great extent by the action of the piston rings.2 Abrasive wear of the piston skirt
When abrasive particles (soot. The chemical loads comprise dilution by fuel. severe abrasive wear scars may form on the piston skirt surface. Piston skirt wear is in certain cases suppressed by a harder coating on the skirt. 2000). where the thermal. When the bore polishing has evolved to a stage of heavy polishing. owing to the distribution of the thrust forces during the different cycles of the engine. acidic combustion products and water vapour from the combustion process. is the first occurrence of wear in a cylinder liner. dust or metallic wear particles) enter the sliding interface between the piston skirt and the cylinder liner. medium and heavy polishing. and most of the oil-retaining honing pattern has been erased. and the removal of oil from the liner surface. The wear of the cylinder liner is higher on the anti-thrust side than on the thrust side of the liner.
76
. 1995)..9. High wear of the cylinder liner is furthermore associated with the top reversal point of the second piston ring. (Dong et al.. The thermal loads cause lubricant degradation by ageing and partial evaporation. and possibly by dust from the intake air that can contribute by causing abrasive wear.3. Priest and Taylor. 1996). A light degree of bore polishing increases the oil consumption. erosive and abrasive conditions are the severest. The wear of the cylinder liner is additionally accelerated by solid carbon particles from the combustion process. Practical observations and theoretical analyses correlate well in terms of the strongest wear of the cylinder liners taking place in the vicinity of the top reversal point of the top piston ring. The wear is normally accompanied by similar abrasive wear on the cylinder liner surface.

albeit under dry running conditions.5 hours (Ma et al. 2000). Owing to the tests being performed without lubrication. 2000). 1996. adhesion. 1998).. Means to suppress the wear of the cylinder liner are presented in Chapters 4 and 10. see Chapter 10) representing certain proportions of the piston ring face surface area (Pawlus. 1997. Sherrington and Mercer. the wear can be estimated from changes in relevant surface roughness parameters or the bearing area curve (the Abbot curve. the wear volume of cylinder liners can be determined by comparison of surface roughness profiles or cross-section profiles before and after the tests (Ohlsson.
77
.. According to tests by Ma and co-workers. Priest and Taylor.. and that the strongest wear occurs close to the top ring reversal point (Henein et al.. the findings are primarily representative of strongly starved sliding conditions. For low wear rates. delamination and ploughing (abrasive two-body wear) can be present in the wear of the cylinder liners (Terheci et al. Kumar et al. 1995). 1998). determined the wear mechanisms responsible for the degradation of cast-iron cylinder liners. 2000. the cylinder liner wear reaches a steady-state after 3. Alternatively. 1998). with wear rates during the first hour of operation that are approximately 12 times those during the second or third hour of operation.. The investigation showed that truncation (plastic deformation of surface profile asperities). Terheki and co-authors have experimentally.Break-in wear tests have shown that the wear of a cylinder liner is predominant at the beginning of the break-in process (Ma et al.

or single phase or composite ceramic powders are used. and the temperature increase owing to thermal energy dissipated at the collision between the powder particles and the piston ring surface. and the use of inlays or semi-inlays covering only part of the ring face is used (ISO 6621-4). 1994). For thermally sprayed coatings. piston ring and cylinder liner system. The coating is built up from a powder. most often cemented carbide powders with a metallic binder phase. widely used surface treatment method for piston rings is surface nitriding.10. hence precautions like edge radiusing (before or after the coating deposition). The temperature of the powder. Higher energy processes include the plasma spraying. cause the particle to partly melt and adhere to the piston ring surface. A third. coatings and surface treatments are commonly applied on one or several components of the piston.
78
. soft coatings for friction reduction are applied. New types of coatings are being developed and it can be forecasted that they will eventually be applied in production engines. Coated ring faces may be susceptible to edge flaking. namely thermal spraying and galvanic coating. the detonation gun spraying and the high-velocity oxy-fuel (HVOF) process (Holmberg and Matthews. which is heated and bombarded onto the piston ring surface. 1998).1 Coating deposition and surface treatment techniques
Functional coatings and surface treatments offer several possibilities to improve the sliding properties of metal surfaces. In addition to this. and even more methods are being developed and experimentally evaluated. Surface technology
10. Surface coating and treatment techniques are continuously developing and new groups of coatings are introduced on the market. In internal combustion engines. The exploded wire spray method is in use for the deposition of FeMo coatings (Durga et al. which is a combined hardening and coating process for producing a graded iron nitride structure on the ring surface region. Thermal spraying Plasma spraying and other types of thermally sprayed coatings are mainly applied on piston rings for large-bore diesel engines.. Two different groups of coating application techniques are currently being commercially used for the deposition of wear-resistant coatings on piston rings. The range of coatings and surface treatments currently used in engine components in the reciprocating system covers a variety of different coating compositions and deposition techniques.

An alternative for improving the tribological performance of aluminium block engines is to use coated pistons (see below).. when comparing surfaces with equal surface roughness.. Brauers and Neuhäuser. 1993).e. partly due to wear protection obtained by a tribofilm formed from wear debris (Ahn et al. Work by Durga and co-authors show a reduction in the coefficient of friction when applying porous plasmasprayed FFS (Stainless steel + Ni-BN) or M-1P (Fe-FeO-C) coatings. with molybdenum coating on the full width of the face or as an inlay covering part of the ring face width.. and even lower coefficients of friction with a porous and honed surface (Durga et al. forming lubricious iron oxides like FeO and Fe3O4. low-heat rejection diesel engines. 1992). are used by many diesel engine manufacturers (ISO 6621-4). 1992). sulphur resistant iron-based alloys containing chromium and molybdenum (Harrison.. HVOF-sprayed chromium carbide coating is suggested by Shuster and co-workers as an efficient protection against scuffing and wear (Shuster et al. which may occur due to blow-by of combustion gases from the combustion chamber (Affenzeller and Gläser. For similar applications. Molybdenum is a tribomaterial used particularly in high-temperature applications (Glaeser. 1989). which can be determined by the standard test ASTM D633. i. 1996. and with non-corroding. 1996). good results have been obtained with Fe-C alloys that partly oxidize during the deposition.
79
. 1997). Gaydos and co-workers have successfully studied the use of a chromia (Cr2O3) cylinder liner coating against a tungsten diselenide (WSe2) piston ring coating (Glaeser and Gaydos. For the same kind of applications. Harrison describes atmospheric plasma-sprayed coatings for cylinder surfaces in series production aluminium engine blocks. 2001). 1998). The resistance of the molybdenum coating against adhesive and cohesive failures. is controlled by the coating deposition process parameters (Babu et al. promising results have been obtained with thermally sprayed titania (TiO2 or TiOx) coatings on aluminium substrates (Buchmann and Gadow. Haselkorn and Kelley have studied the tribological performance of ring and liner specimens coated with plasma-sprayed high-carbon-ironmolybdenum and plasma-sprayed chromia-silica coatings. 2002). In a recent overview. 1999). For adiabatic diesel engines. and found them of potential use on cylinder liner or piston skirt sliding surfaces (Haselkorn and Kelley.Molybdenum coated piston rings. Experimental investigations by Ahn and co-workers on plasma-sprayed zirconia thermal barrier coatings against chromiumplated steel in unlubricated reciprocating motion at 200°C specimen temperature have shown that the wear resistance of the coatings was better than that of cast iron in the same kind of tests. Coating by molybdenum is especially useful for preventing hot gas erosion deterioration of the piston ring face surfaces. Engine tests with different advanced chromium carbide (Cr2C3) based sprayed coatings have shown the potential of this group of coatings.

2001). Precautions in the deposition process or post treatment may be necessary for avoiding hydrogen embrittlement of the coated surface material. and Cr2O3-based coatings on cast-iron rings and cylinder liners. 2000). the components to be coated act as cathodes. 1997). Full-scale engine testing of aluminium bronze coated cast-iron piston rings against Cr3C2 based coatings on cylinder liners. all hard chromium-plated piston rings are lapped or polished after the coating deposition. 1997. acts as the chromium donor. honed. 1994). Mollenhauer. 1998. The most widely used galvanic coating for the wear protection of piston rings is hard chromium plating (Affenzeller and Gläser. From an environmental protection point of view. lapped or polished before the surfaces are taken into use. Plasma-sprayed or otherwise thermally sprayed layers on the face surface of a piston ring can be applied as a uniform (with minor deviations) layer or as an inlay in a groove. 1996). A feature typical of hard chromium coatings on piston rings is the network of minute cracks that covers the surface and provides pockets for lubricating oil. chromic acid (H2CrO4). an electrolyte and an electric power source. As the as-coated hard chromium surface is rather coarse. Electrochemical coating Galvanic. grinding or polishing. whereas an inert anode is used (Newby. 2000). metal ions from the anode move through the electrolyte to the substrate at a rate that is defined by the electric current density (Holmberg and Matthews. In the coating process. followed by cleaning before being taken into use (Newby. Different mechanical and chemical techniques are applied for obtaining the porous topography. In electrochemical plating. In practice. 2001). and the surface finishing needs to be optimised individually for each coated system (Radil. In electrochemical chromium plating. lapping. have given promising results in terms of component wear and lubricant sensitivity (Miyake et al. ISO 6621-4). or the aqueous solution of chromium trioxide (CrO3). it normally requires honing. Originally the porous chromium coating was developed for piston rings and aluminium cylinder bores of aircraft and diesel engines. and locations for hydrogen gas formation. the use of chromium plating on
80
.. Since the 1940´s hard chromium coatings have often been produced as porous.particularly when molybdenum is included in the coating composition (Rastegar and Richardson. Thermally sprayed ceramic or ceramic-metal composites need to be ground. which means that the effective width of the coated part is less than the actual width of the ring face (Federal Mogul. coatings are based on a donor anode. or electrochemical. and during this post-treatment the ring can simultaneously be given its barrel or other profile shape.

610°C depending on process). 1999). Severe wear of Cr-plated piston rings may. temperature and time. 1996).piston rings is less desirable than the use of non-chromium-plated steel rings (Affenzeller and Gläser. Chromium-plated piston rings have proven their durability in long-term use. 10 minutes to 4 hours. Gas nitriding is carried out in ammonia gas. Improvements in the wear and scuffing resistance of aluminium pistons in overeutectic aluminium cylinder liners is obtained by ferritic iron plating of the piston skirt (Affenzeller and Gläser. 1996). The thickness of the nitride layer and the thickness of the diffusion zone depend on the process and the substrate
81
. 20. with the addition of carbon dioxide or carbon monoxide gas for carbonitriding. Plasma nitriding uses a mixture of oxygen and nitrogen gas as the nitrogen donor. 1998).. Scuffing experiments by Wang and Tung with various coatings on piston skirts in aluminium cylinder liners have shown that a selectively plated Ni-W coating and suspended ceramic particulate plated Ni-P-BN can provide the same tribological benefits as an iron coating (Wang and Tung.. 1995). however. while molten salt nitriding gives the shortest time.. Hard chromium coatings are occasionally used for improving the wear resistance of ring grooves in steel composite piston crowns (Röhrle. at the outermost region of the piston ring material. The reaction kinetics. Gas nitriding requires the longest processing time. a molten salt bath or a plasma at high temperature (450. and a diffusion zone between the nitride layer and the substrate material. the depth of diffusion of the nitrogen atoms into the steel structure and the hardness depth are controlled by the process. occur owing to the lack of oil retention volume on the cylinder liner surface. or thermomechanical fatigue of the coating (Shuster et al. cyanides and cyanates are normally used. and the process is activated by electrical discharges in an insulated chamber. and reactions between the iron. the steel component is exposed to a nitrogencontaining atmosphere. The abrasive wear resistance of a hard chromed piston ring can be improved by adding hard ceramic particles to the coating during the deposition of the coating (Federal Mogul. 1998)... The process is terminated by cooling at a rate that gives the desired hardness and toughness properties. 1999).100 hours. Nitriding and carbonitriding Surface nitriding or carbonitriding are processes for producing a hardened surface consisting of an iron nitride. During the surface treatment. nitrogen and carbon atoms. Galvanic (electrochemical) coating materials furthermore include copper (Cu) and tin (Sn) as friction-reducing coatings for piston ring face surfaces (Federal Mogul. martensitic transformation on the cylinder liner surface. In molten salt bath processes. which causes diffusion of nitrogen (and carbon) into the steel. or a harder iron carbo nitride.

In engine tests. partly due to the formation of aluminium nitride (AlN) in the surface structure (Bindumadhavan et al. as for instance oxygen-sulphur nitriding of grey cast iron has been recently studied (Baranowska. traditional chromium-plated piston rings turned out to be equally or more wear-resistant than carbonitrided piston rings in engine tests. with 1 300 HV as an average typical of various steel and cast-iron piston ring materials (Brauers and Neuhäuser. Graphite coatings Graphite coatings are deposited on aluminium and cast iron piston skirt surfaces as a resin that contains pigments. for which chromiumplated or molybdenum spray-coated rings are not required. 1989. Corrosive attacks on the carbonitrided rings were obvious. Ferro-oxidising Ferro-oxidising can be carried out on all surfaces of an otherwise uncoated piston ring (ISO 6621-4). In most cases. The nitriding process is subject to development efforts. For smaller internal combustion engines operating under lower loads. Low-carbon steel surfaces that are aluminised before being nitrided reveal a higher hardness than nitrided plain low-carbon steel. 1989). boron nitride. the system wear of the ring flank and the ring groove were equal or larger than in the case of untreated cast iron rings.material. irrespective of the nitriding process applied. who
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.. where carbonitrided chromium-alloyed steels were superior to carbonitrided cast iron or low-alloy steels in terms of wear resistance. and their role in the total wear of the rings may be significant. as has been the case in certain production engines. The adhesion of the graphite coating can be improved by applying a metallic phosphate layer onto the surface prior to the application of the coating (Röhrle. particularly in engines that are frequently used for shorter duration (Brauers and Neuhäuser. 2000). carbonitrided rings can be used. Mollenhauer. 1998). 1998). Federal Mogul. 1995. The surface hardness obtained by the nitriding process and the hardness gradient in the surface region depend on the material. An atomised spray coating consisting of epoxy resin. molybdenum disulphide and graphite on the piston skirts and cylinder liners has been studied by Durga and co-workers. The response of different piston ring materials to carbonitriding is obvious from the differences in their wear rates in engine tests. The wear of carbonitrided piston rings made from a specific steel has been found almost equal in high-temperature tests. 1997). A method called Grafal® is based on a phenolic singlestage resin with fine colloidal graphite.

Phosphate coatings A phosphate (zinc phosphate or manganese phosphate) coating can be formed onto piston rings by chemical reaction with phosphate crystals. Ion exchange reactions Thin layers of lead (Pb) and tin (Sn) are produced onto as-machined piston skirt surfaces by ion-exchange reactions. Thicker layers are obtained by a hard anodising process. 1995). 1995). Anodising Anodising is the name given to for a controlled electrochemical process for obtaining a dense aluminium oxide grading on the surface of an aluminium component. 1996). Ion implantation increases the hardness of metals without significantly affecting the microgeometry of the surface. by applying Pb and Sn salts onto the surfaces for a certain reaction time. Ion implantation A procedure to expose a metal surface to accelerated metal ions. ISO 6621-4). The phosphate layer is softer than the piston ring material and supports effective running-in. and the phosphate treatment can reduce the formation of wear scars (Federal Mogul. and they are commercially widely used (Röhrle. which has been proven successful in several applications. titanium nitride and chromium
83
. which diffuse into a certain depth of the material. have. is called ion implantation. shown by pin-on-disc and reciprocating tests that nitrogen ion implantation alone is insufficient for providing improvements in the tribological performance of aluminium piston materials (Qiu et al.. Studies by Qui et al. however. Hard anodising is occasionally applied on piston tops for thermal protection reasons (Röhrle. 1998). Anodising has been studied with reference to its use on piston skirts against aluminium cylinder liners. The thin Pb and Sn layers are soft and improve the running properties at the piston skirt and cylinder liner contact. 1998.have found that the amount of boron nitride must be kept below a certain limit in order to suppress cylinder liner wear (Durga et al. like nitrogen or titanium. Vacuum methods The novel surface coating techniques that have potential for piston rings can be found among the group of hard diamond-like carbon coatings. but turned out to be a less successful solution (Wang and Tung.. 1999).

.
84
.. This group of coatings offers potential for use as coatings on piston rings. particularly when the DLC coating is doped with appropriate metal ions (Arps et al. for reducing the wear and the friction.titanium nitride (Ti/TiN) compound coatings in comparison with electroplated Cr coatings and phosphated ring surfaces have been demonstrated by Zhuo and co-workers by model experiments with piston ring segments and cylinder liner segments. The implementation of the latter results into internal combustion engines will require that the results of complementary investigations are taken into consideration.nitride coatings and other types of PVD and CVD coatings produced under partial vacuum.. have been shown by Vetter and Nevoigt (Vetter and Nevoigt. who have found that these coatings reduce the wear to about one-tenth of the wear of electroplated chromium. By applying different coating deposition parameters. for instance. which all revealed less wear when coated with the Ti/TiN coatings (Zhuo et al. 2000). 2001). 1996). 1999). while the coefficient of friction was similar for the CrxN and the Cr coatings. 1999). 1997. the amount of. The necessity to redesign the entire tribosystem with reference to the thin hard coatings is pointed out by the authors (Friedrich et al. 1999). The tribological benefits of applying a metalcontaining a-C:HMe hard carbon coating on pistons for hydraulic cylinders. sp2 (graphite) and sp3 (diamond) bonds and hydrogen in the DLC structure can be controlled. Miniature tests with DLC-coated piston ring samples in an oscillating rig have shown that this group of coatings can offer wear and friction reductions. Haselkorn and Kelley have studied the tribological performance of ring and liner specimens coated with lowtemperature arc vapour deposited CrN coatings and found the coatings of potential use on cylinder liner or piston skirt sliding surfaces (Haselkorn and Kelley. Scuffing experiments by Wang and Tung with various coatings on piston skirts for aluminium cylinder liners have shown that a PVD DLC coating with a silicon interlayer can provide lower friction but causes more wear than current iron coatings (Wang and Tung. 1992).. which strongly affects the tribological properties of the coating. Friedrich and co-workers. The benefits of multi-layered titanium . Chromium nitride (CrN and Cr2N) coatings on piston rings have been studied in model wear tests by Broszeit. The diamond-like hard carbon coatings (DLC) have shown beneficial friction and wear properties in tribological investigations (Ronkainen. Broszeit et al.

surface waviness and surface roughness of pistons and cylinder liners (Lenhof and Zwein. Rk or Sk. and new parameters are continuously being taken into use. are statistically representative values with the following definitions (compare with Fig.1 Surface roughness basics
Traditionally.2 Surface roughness
Deviations from absolute mathematical geometry can be expressed as form deviations. the primary need is to determine the surface roughness. ring and cylinder liner surfaces.
10.
85
. piston skirts and cylinder liners has been taken into particular consideration in the development of surface roughness measurement techniques and the presentation of measurement results as surface roughness parameters. Form errors. The different aspects refer to the magnitude of the wavelength of the deviations from absolute form. Shuster et al. certain surface texture height parameters were developed some decades ago. or the long-term running surface that determines the life time of the component. The surface roughness of piston rings.10. resulting in 2-dimensional surface roughness parameters. 1999): Rpk or Spk. More developed equipment nowadays offers an option to perform surface roughness measurements on a surface. surface waviness and surface roughness. For distinguishing between 2dimensional and 3-dimensional surface roughness parameters. the former ones are denominated Rsubscript and the latter ones Ssubscript. 1996.2. The surface texture height parameters. the surface roughness measurements have been carried out along lines. reduced peak height. or the surface profile peaks that are erased by running-in. 10.1) (Ohlsson. 2002) can be metrologically determined from roundness and surface roughness measurements. which are parameter presentations of the information in bearing area curve (the Abbot curve). for obtaining 3-dimensional surface roughness parameters. core roughness depth. For piston rings. with form errors representing the longest wavelength and surface roughness the shortest wavelengths of the deviation (Ohlsson. Each of the above parameters has an influence on the tribological performance of the piston/cylinder couple. 1996). which are elastic.. For piston. Surface roughness parameters for common use are the arithmetic average surface roughness (Ra or Sa) and the profile skewness (Rsk or Ssk).

and which indicates how the surface irregularities are repeated on the surface.2. which can be expressed by means of surface roughness parameters. MR1 (or Sr1) and MR2 (or Sr2). Rvk/Rk. The mechanically textured cylinder liner pattern. Graphical representation of the surface texture height parameters for a 2dimensional measurement. is of fundamental importance to the tribological performance of. this ratio expresses the plateauness of the cylinder liner texture. 10. the so-called plateau-honed cylinder. for instance. 1996). is a good example of improved tribological behaviour by surface texturing. or the oil retaining capability of the groove pattern of a honed surface or similar. chromium-plated piston rings can be lapped subsequently to the deposition of the coating. a plateau surface has a Rvk/Rk ratio of about three. reduced trough depth. The texture formed by a network of oil channels on certain type of chromium-plated piston rings is similarly beneficial as the honing pattern on a cylinder
86
. To some extent each method and type of equipment gives is characteristic features to the surface roughness parameters. the material components (in %) determined for the line of intersection coinciding with the upper and lower limits of the roughness core profile. as well as the texture. a piston ring and cylinder liner contact. The surface roughness can be determined by a variety of methods and equipment.1. For the same reason.Rvk or Svk.
Rvk
Rk
Rpk
0% MR1
MR2
100%
Fig.2 Topography and texture
The magnitude of the surface roughness. which means that the results are only partly intercomparable (Ohlsson.
10.

1999a..
10. 1998). Recent experiments by Galligan and co-workers have again shown the role of an optimal cross-honed texture with oil retention grooves on the cylinder liner for the prevention of piston ring scuffing (Galligan et al.2. 1999b). 2001). 2001.liner (Federal-Mogul. Steinhoff et al.. which allows the production of any pattern of cavities onto a piston ring or other sliding surface (Ronen et al..3 Textured surfaces
Post-processing of coated (or uncoated) piston ring surfaces has recently been developed by the introduction of the laser texturing technique.
87
.

the tightening exhaust emission legislation causes a general pressure to reduce the oil film thickness in the ring/liner contact.11. surface quality and texture in terms of honing. The thinnest oil films occur in the vicinity of the dead centres. disturb the oil film formation in the ring/liner contact. rings and cylinder liner. ring radial force and transients.
88
. and is consequently subjected to gas pressure differences with strong variations. carbon deposits and particles originating from the combustion process contribute to the wear of the piston. for ring clearance reasons and for maintaining the gas loading of the rings. The spring loads of the rings are responsible for the rest of the contact pressure. The blow-by leakage may. however. sliding velocity and transients. a large part of which is conducted to the cylinder liner surface by the piston ring pack. and under unfavourable conditions the blow-by may rise to a level at which a hot gas erosion damage occurs on the piston ring. As oil entrains into the exhaust gases and thus increases the emissions. acidic combustion products. by allowing the cylinder pressure to act on the back-side of the rings. The gas pressure gradient across the piston is utilised for increasing the contact pressure between the first. an oil composition with poor lubricating properties is formed on the piston rings. The heat produced in the combustion of the fuel. and the contact geometry and its variations due to ring twist. The piston acts as a seal between the combustion chamber and the crankcase. the oil transport. When the combustion products and wear particles from the piston ring area are mixed into the lubricating oil on the piston rings. axial and radial motion of the ring and cylinder bore deformation. in particular the TDC. The oil ring is normally loaded by a stronger spring force and operates under thinner oil films than the compression rings on average. Limited blow-by leakage of combustion gas from the combustion chamber through the piston ring pack to the crankcase is allowed. surface materials. interaction of nearby rings. For advanced computer simulation of the lubrication conditions of the ring/liner contact. Water vapour. Summary
The piston of an internal combustion engine is the first component of the mechanism converting the chemical energy of the fuel into mechanical work. reduces the hardness and wear resistance of the piston and ring materials and causes oxidation and evaporation of the oil on the upper cylinder walls. compression ring against the cylinder liner surface. most of the above parameters can nowadays be taken into consideration. Measurements of the oil thickness in the ring/liner contact provide a valuable tool for the tribological analyses of the ring-pack lubrication conditions. with the interactions of the complete piston ring pack and oil starvation issues as recent and highly relevant features. Lubrication conditions at ring/liner contacts strongly depend on the oil quality. and partly the second.

where the sliding velocity is at its lowest. for example as the result of cylinder bore polishing that has erased the honing pattern from the cylinder liner. Under misfortunate conditions. Hard piston ring coatings. particularly at TDC locations where the combustion pressure rises to its maximum level. indicate the probability for wear of the sliding surfaces. In the vicinity of the reversal points of the piston motion. with the mission of reducing the friction under conditions of boundary lubrication. like chromium plating. Severely contaminated lubricating oil has the potential to cause three-body abrasive wear of the sliding counter surfaces. thermally sprayed molybdenum or surface nitriding. are currently applied for protecting the piston rings against abrasive wear. as indicated by locations of intensive cylinder liner wear.
89
. adhesive wear between ring and cylinder liner may occur and result in piston ring scuffing. and reflect the lubrication conditions at the respective position of the piston. Most of the wear takes place in the vicinity of the top dead centre and the bottom dead centre of the piston motion. Owing to the above preconditions. the wear of the sliding surfaces of a piston assembly and a cylinder liner sets to a steady level of mild wear. as the result of surface asperity interactions in the sliding interface. damage by combustion gas erosion and fatigue damages. After an initial phase of running-in. are applied onto the piston skirts and occasionally onto rings for reducing friction during the running-in stage.The coefficient of friction and the friction force between the respective components of the piston assembly against the cylinder liner reflect the energy consumption of the oscillating components. The wear mechanism responsible for the running-in wear is a two-body abrasion. the coefficient of friction reaches its maximum values. the ring and liner contact operates under conditions of boundary or mixed lubrication in the vicinity of the dead centres and hydrodynamic lubrication at the mid-stroke regions of the piston motion. Soft coatings.

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VTT Information Service P.O. not least indicated by the large number of articles published on this topic in recent years. wear. The work is intended as a compact reference volume for internal combustion engines in general. +358 9 456 4404 Fax +358 9 456 4374
. P. with particular emphasis on diesel engines. the wear of the sliding surfaces and surface technology for wear reduction.
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National Technology Agency Tekes. Jaana & Sandström. the sealing action. Metallimiehenkuja 6. This literature survey. Central topics discussed in this work are the basic functions of the piston and the piston rings. Peter.Published by Series title. Finland Phone internat.inf. mechanical and thermal loads on the rings. Finland
ISBN Project number
951–38– 6107–4 (soft back ed.O. FIN–02044 VTT. exhaust emissions. FIN–02044 VTT. Recent studies include modelling. miniaturised experimental work and full-scale engine testing. the design and the materials of the components.Box 1702. tribology
Activity unit
VTT Industrial Systems. Tamminen. covering over 150 references.
Keywords
Piston ring. the contact pressure between ring and liner. number and report code of publication
VTT Research Notes 2178 VTT–TIED–2178
Author(s)
Andersson.) 951–38–6108–2 (URL: http://www. the lubrication conditions and the influence of combustion products. Carl-Erik
Title
Piston ring tribology
A literature survey
Abstract
The tribological considerations in the contacts formed by the piston skirt.inf. piston rings and cylinder liner have attracted much attention over several decades.fi/pdf/)
Date Language Pages
V9SU00507
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December 2002
Name of project
English
105 p. Fortum Oil & Gas Oy.Box 2000. Sisu Diesel Oy. the coefficient of friction and the friction force. VTT
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VTT Tiedotteita – Research Notes 1235–0605 (soft back edition) 1455–0865 (URL: http://www.vtt. aims to shed new light on the tribological issues related to the piston assembly. friction. blow-by leakage. Volvo Technological Development Corporation. Wärtsilä Corporation. lubrication.vtt. hot gas erosion damage.

the contact pressure between ring and liner.
Piston ring tribology.VTT RESEARCH NOTES 2178
The tribological considerations in the contacts formed by the piston skirt.) ISSN 1235–0605 (soft back ed. Central topics discussed in this work are the basic functions of the piston and the piston rings. miniaturised experimental work and full-scale engine testing. the sealing action.vtt.)
ISBN 951–38–6108–2 (URL: http://www.fi/pdf/) ISSN 1455–0865 (URL: http://www. (09) 456 4404 Fax (09) 456 4374
This publication is available from VTT INFORMATION SERVICE P. hot gas erosion damage. the design and the materials of the components.inf. Recent studies include modelling. aims to shed new light on the tribological issues related to the piston assembly. not least indicated by the large number of articles published on this topic in recent years.inf. + 358 9 456 4404 Fax + 358 9 456 4374
ISBN 951–38–6107–4 (soft back ed. the coefficient of friction and the friction force. mechanical and thermal loads on the rings. This literature survey.O.Box 2000 FIN–02044 VTT. The work is intended as a compact reference volume for internal combustion engines in general.vtt. (09) 456 4404 Faksi (09) 456 4374
Denna publikation säljs av VTT INFORMATIONSTJÄNST PB 2000 02044 VTT Tel. piston rings and cylinder liner have attracted much attention over several decades. blow-by leakage. covering over 150 references. exhaust emissions. the lubrication conditions and the influence of combustion products. with particular emphasis on diesel engines. the wear of the sliding. Finland Phone internat. A literature survey
Tätä julkaisua myy VTT TIETOPALVELU PL 2000 02044 VTT Puh.fi/pdf/)
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